correlation based thermal design of air transport
TRANSCRIPT
CORRELATION BASED THERMAL DESIGN OF AIR TRANSPORT RACK CHASSIS
A THESIS SUBMITTED TO THE GRADUATE SCHOOL OF NATURAL AND APPLIED SCIENCES
OF MIDDLE EAST TECHNICAL UNIVERSITY
BY
BEKĐR ONUR ÇOLPA
IN PARTIAL FULFILLMENT OF THE REQUIREMENTS
FOR THE DEGREE OF MASTER OF SCIENCE
IN MECHANICAL ENGINEERING
AUGUST 2011
Approval of the thesis:
CORRELATION BASED THERMAL DESIGN OF AIR TRANSPORT RACK CHASSIS
submitted by BEKĐR ONUR ÇOLPA in partial fulfillment of the requirements for the degree of Master of Science in Mechanical Engineering Department, Middle
East Technical University by, Prof. Dr. Canan Özgen ________________ Dean, Graduate School of Natural and Applied Sciences
Prof. Dr. Süha Oral ________________ Head of Department, Mechanical Engineering
Assoc. Prof. Dr. Đlker Tarı ________________ Supervisor, Mechanical Engineering Dept., METU
Examining Committee Members:
Asst. Prof. Dr. Cüneyt Sert ________________ Supervisor, Mechanical Engineering Dept., METU
Assoc. Prof. Dr. Đlker Tarı ________________ Supervisor, Mechanical Engineering Dept., METU
Asst. Prof. Dr. Ahmet Yozgatlıgil ________________ Supervisor, Mechanical Engineering Dept., METU
Asst. Prof. Dr. Tuba Okutucu Özyurt ________________ Supervisor, Mechanical Engineering Dept., METU
Mustafa Ocak, M.Sc. ________________ Mechanical Engineer, ASELSAN
Date: 19.08.2011
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I hereby declare that all information in this document has been obtained and
presented in accordance with academic rules and ethical conduct. I also declare
that, as required by these rules and conduct, I have fully cited and referenced
all material and results that are not original to this work.
Name, Last Name : Bekir Onur ÇOLPA
Signature :
iv
ABSTRACT
CORRELATION BASED THERMAL DESIGN OF AIR
TRANSPORT RACK CHASSIS
Çolpa, Bekir Onur
M.Sc., Department of Mechanical Engineering
Supervisor: Assoc. Prof. Dr. Đlker Tarı
August 2011, 117 pages
In this thesis, a Thermal Model Tool (TMT) is developed for standard Avionic
Transport Rack (ATR) chassis and thermal design of a standard ATR chassis is done
using developed TMT. This ATR chassis is a Digital Moving Map (DMAP) of a
helicopter and the tool is used to determine the cooling channel details of DMAP.
TMT decreases design process steps and eliminates the complexity of the design.
Experimental studies are conducted on one of the existing chassis produced in
Aselsan Inc. for different operating conditions. There are two different operating
conditions for the chassis as 25 ºC and 55 ºC, which are given, in military standard
MIL-STD-810F. Critical temperature values are measured, which are used in
analytical calculations, and results are represented.
At the first step, outputs of the experimental studies are used in analytical calculation
in order to develop TMT. Secondly, heat dissipation rate of two different chassis are
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calculated easily by using the TMT, and without making effort for CFD analysis, the
necessary number of plate fins of the chassis are assessed considering given
geometrical constraints and heat loads. Finally, cooling channels are generated using
the results of TMT.
In the next step the chassis, which are designed using the results of TMT, are
analyzed numerically by using Icepak Computational Fluid Dynamics (CFD) tool
and results of TMT are verified. The cooling capacities of the decided plate fins,
which are obtained by TMT, are checked whether or not the required heat dissipation
rates are ensured.
Consequently, TMT is tested under for two different operating conditions on two
different chassis. Analytical and numerical studies for both conditions are compared
and discussed in detail. Comparisons show that, developed TMT results are
meaningful and close to numerical results, therefore TMT can be used in
forthcoming ATR chassis designs.
Keywords: ATR chassis, Forced Convection Cooling, CFD.
vi
ÖZ
ATR STANDARTLARINDAKĐ ŞASENĐN KOLERASYON
TABANLI TERMAL TASARIMI
Çolpa, Bekir Onur
Yüksek Lisans, Makina Mühendisliği Bölümü
Tez Yöneticisi: Doç. Dr. Đlker Tarı
Ağustos 2011, 117 sayfa
Bu tez çalışmasında standart Avionic Transport Rack (ATR) şaseler için Termal
Modelleme Aracı (TMA) geliştirilmiş ve bu araç kullanılarak standart bir ATR
şasenin termal tasarımı yapılmıştır. Bu ATR şase bir helikopterin Hareketli Sayısal
Harita Birimidir (HSHB). TMA HSHB’nin soğutma kanallarının detaylarının
belirlenmesinde kullanılmaktadır. TMA tasarım sürecini kısaltmakta ve tasarım
aşamasındaki karmaşıklıkları ortadan kaldırmaktadır.
Aselsan A.Ş. bünyesinde mevcut olan bir şase üzerinde 25 ºC ve 55 ºC ortam
koşullarında deneyler yapılmıştır. Bu ortam koşulları askeri standart olan MIL-STD-
810F dokümanında belirtilmektedir. Deneyler sırasında, analitik çalışmalarda
kullanılmak üzere kritik olan sıcaklık değerleri ölçülmüş olup sonuçlar verilmiştir.
Deneysel sonuçlar TMA’nın geliştirilmesi için kullanılmıştır. Herhangi bir
Hesaplamalı Akışkanlar Dinamiği (HAD) analizi yapılmadan TMA kullanılarak iki
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faklı şase için gerekli olan plaka kanatcık sayıları hesaplanmıştır. Bu hesaplama
işlemi sırasında önceden verilmiş olan ısı yükleri ve geometrik kısıtlamalar göz
önünde bulundurulmuştur. TMA sonuçlarına göre şaselerin soğutma kanalları
şekillendirilmiştir.
Daha sonra, soğutma kanal detayları belirlenmiş olan şase HAD aracı olan Icepak
kullanılarak analiz edilmiştir. Böylece, karar verilen plaka kanatçıkların ısı atım
kapasitelerinin doğruluğu kontrol edilmiştir.
Sonuç olarak, TMA iki farklı çalışma koşulu için iki farklı şase üzerinde test
edilmiştir. Analitik ve numerik çalışmalar her iki çalışma şartı için de karşılaştırılmış
ve detaylı olarak incelenmiştir. Değerlendirmeler neticesinde TMA sonuçlarının
anlamlı ve numerik çalışma sonuçlarına yakın olduğu ortaya çıkmıştır. TMA’nın
yeni ATR şase tasarımları sırasında kullanılabileceği görülmüştür.
Anahtar Kelimeler: ATR şaseler, Zorlanmış Taşınımla Soğutma, HAD.
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To My Wife
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ACKNOWLEDGMENTS
The author wishes to express his sincere appreciation to his supervisor Assoc. Prof.
Dr. Đlker TARI for the endless support, encouragement, and patience during his
research activities.
The author would like to thank ASELSAN, Inc. and his manager Mr. Đhsan ÖZSOY,
for his support and guidance in his study and let to use experimental facilities of
mechanical/optical design department.
The author is thankful to his superior Serkan DÖRTKARDEŞLER for the
encouragement and guidance. Without his unconditional support in guiding the
author to study electronics cooling, this dissertation would never have come into
existence.
The author would like to express his appreciation to his colleagues Mustafa OCAK
and Ahmet Hakan SEZGĐN for their assistance and valuable support.
Lastly, the author would like to express his endless gratitude to his wife for her love,
support, and faith in him.
x
TABLE OF CONTENTS
ABSTRACT................................................................................................................ iv
ÖZ ............................................................................................................................... vi
ACKNOWLEDGMENTS .......................................................................................... ix
TABLE OF CONTENTS............................................................................................. x
LIST OF TABLES .....................................................................................................xii
LIST OF FIGURES .................................................................................................. xiv
LIST OF SYMBOLS ...............................................................................................xvii
CHAPTERS ................................................................................................................. 1
1 INTRODUCTION..................................................................................................... 1
1.1 ATR Chassis....................................................................................................... 3
1.1.1 Flow-Through Cooling Method .................................................................. 5
1.1.2 Natural Convection Cooling Method.......................................................... 6
1.2 VME Cards......................................................................................................... 9
1.3 Vaneaxial Fans ................................................................................................. 13
1.4 Literature Review............................................................................................. 14
2 EXPERIMENTAL STUDIES................................................................................. 26
2.1 Assumptions..................................................................................................... 29
2.1.1 Assumption-1 ............................................................................................ 29
2.1.2 Assumption-2 ............................................................................................ 30
2.2 25 °C Ambient Temperature Experiments....................................................... 34
2.3 55 °C Ambient Temperature Experiments....................................................... 39
3 ANALYTICAL STUDIES...................................................................................... 43
3.1 DMAP Analytical Studies................................................................................ 44
3.1.1 Heat Transfer Calculations........................................................................ 47
3.1.2 Pressure Drop Calculations....................................................................... 50
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3.1.3 Thermal Model Generation and Solutions ................................................ 53
3.1.3.1 25 ºC Operating Condition Calculations............................................ 54
3.1.3.2 55 ºC Operating Condition Calculations............................................ 62
3.1.3.3 25 ºC and 55 ºC Fin Number Comparison ......................................... 63
3.1.4 Radiative Heat Transfer Calculations ....................................................... 65
3.2 ACCC Analytical Studies ................................................................................ 73
4 NUMERICAL STUDIES ....................................................................................... 76
4.1 Numerical Studies of DMAP ........................................................................... 77
4.1.1 Boundary Conditions and Basic Parameters of DMAP ............................ 78
4.1.2 Grid Generation on DMAP ....................................................................... 80
4.1.3 Numerical Solution of DMAP .................................................................. 83
4.1.3.1 25 ºC Operating Condition Solution of DMAP ................................. 84
4.1.3.2 55º Operating Condition Solution of DMAP ..................................... 87
4.2 Numerical Studies of ACCC............................................................................ 88
4.2.1 Boundary Conditions and Basic Parameters of ACCC............................. 90
4.2.2 Grid Generation on ACCC........................................................................ 90
4.2.3 Numerical Solution of ACCC ................................................................... 91
4.2.3.1 25 ºC Operating Condition Solution of ACCC .................................. 91
4.2.3.2 55 ºC Operating Condition Solution of ACCC .................................. 93
5 DISCUSSION ......................................................................................................... 95
REFERENCES........................................................................................................... 99
APENDICES............................................................................................................ 102
A. COOLING CHANNEL DIMENSIONS OF DMAP .......................................... 102
B. FAN PERFORMANCE CURVE AND DIMENSIONS..................................... 104
C. MATHCAD CODE FOR FIN OPTIMIZATION ............................................... 108
D. ANALYTICAL RESULTS................................................................................. 114
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LIST OF TABLES
Table 1.1 Standard ATR case dimensions [1].............................................................. 4
Table 1.2 Power loads of the VME cards. ................................................................. 11
Table 1.3 General Specifications of Aximax 2 [12]. ................................................. 14
Table 1.4 Fin manufacturing limits based on Iyengar [10].................................... 15
Table 1.5. Manufacturability Constraints [11]........................................................... 17
Table 1.6 Thermal performance of heat sink [26]...................................................... 25
Table 2.1 Power loads of cards .................................................................................. 35
Table 2.2 Opposite channels thermocouple results.................................................... 39
Table 2.3 Measured slot temperatures at 25 ºC.......................................................... 39
Table 2.4 Measured slot temperatures at 55 ºC.......................................................... 42
Table 3.1 Fan Curve Data (DMAP). .......................................................................... 50
Table 3.2 Air properties at 25 ºC. .............................................................................. 54
Table 3.3 DMAP channel dimensions. ...................................................................... 55
Table 3.4 Critical temperatures of 25 ºC operating condition.................................... 58
Table 3.5 LTD and efficiency of channel at 25 ºC..................................................... 59
Table 3.6 Calculation results at 25 ºC for the number of plate fins 15. ..................... 60
Table 3.7 Calculation results at 25 ºC for the number of plate fins 17. ..................... 61
Table 3.8 Air properties at 55 ºC. .............................................................................. 62
Table 3.9 Critical temperatures of 55 ºC operating condition.................................... 62
Table 3.10 Calculation results at 55 ºC for the number of plate fins 21. ................... 63
Table 3.11 25 ºC and 55 ºC results comparison. ........................................................ 63
Table 3.12 Wall Numbers. ......................................................................................... 67
Table 3.13 View factors. ............................................................................................ 69
Table 3.14 Fan Curve Data (DMAP). ........................................................................ 74
Table 3.15 ACCC Channel Dimensions. ................................................................... 75
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Table 3.16 Calculation results for 24 number of plate fins (ACCC). ........................ 75
Table 4.1 25 ºC results of DMAP for different number of elements. ........................ 84
Table 4.2 Temperature results of points at 25 ºC operating condition (DMAP). ...... 86
Table 4.3 Results of points at 25 ºC operating condition (DMAP). .......................... 86
Table 4.4 Temperature results of points at 55 ºC operating condition (DMAP). ...... 87
Table 4.5 Temperature results of points at 25 ºC operating condition (DMAP). ...... 88
Table 4.6 Temperature results of points at 25 ºC operating condition (ACCC). ....... 92
Table 4.7 Temperature results of points at 25 ºC operating condition (ACCC). ....... 93
Table 4.8 Temperature results of points at 55 ºC operating condition (ACCC). ....... 93
Table 4.9 Temperature results of points at 25 ºC operating condition....................... 94
Table 5.1 Results of experimental studies. ................................................................ 96
Table 5.2 Analytical and numerical results of DMAP............................................... 96
Table 5.3 Analytical and numerical studies results of ACCC. .................................. 97
xiv
LIST OF FIGURES
Figure 1.1 Example of a standard ATR chassis [3]. .................................................... 2
Figure 1.2 Example of a standard VME card [3]. ........................................................ 2
Figure 1.3 Standard ATR case dimensions [23]. ......................................................... 3
Figure 1.4 Flow-through cooling chassis [2] ............................................................... 6
Figure 1.5 Forced convection cooling chassis with cooling channels [3].................... 7
Figure 1.6 Natural Convection cooling chassis [4]...................................................... 7
Figure 1.7 Draft model of the chassis. ......................................................................... 8
Figure 1.8 Placement of Fan. ....................................................................................... 9
Figure 1.9 Standard VME card sizes [5]. ................................................................... 10
Figure 1.10 Thermal Plate.......................................................................................... 10
Figure 1.11 Placement of slots. .................................................................................. 12
Figure 1.12 Aximax 2 fan [12]................................................................................... 13
Figure 1.13 Fin thickness and number influence on fin stack thermal resistance [9] 16
Figure 1.14 Short chassis, top to bottom flow, 250Pa pressure drop, 85C edge [9]. . 16
Figure 1.15 Side-inlet-side-exit (SISE) rectangular plate fin heat sink conf. [11]. ... 17
Figure 1.16 Practical model of a conventional plate-fin heat sink [20]. .................... 18
Figure 1.17 Rectangular model: T0= 300 K, Tw =423 K [24].................................... 20
Figure 1.18 Aerodynamic model: T0= 300 K, Tw =423 K [24]. ................................ 20
Figure 1.19 Rounded inlet model (Rounded I): T0= 300 K, Tw =423 K. [24]. .......... 20
Figure 1.20 Rounded inlet and outlet model: T0= 300 K, Tw =423 K [24]................ 21
Figure 1.21 The geometric parameters [25]. .............................................................. 22
Figure 1.22 The effect of the stagger when the spacing is fixed at Wopt (β = 1) N=4,
ReL=103) [25]. ............................................................................................................ 23
Figure 1.23 The effect of the stagger parameter on the optimal spacing and the
maximum overall thermal conductance (A = 0.5, N = 4, ReL = 103) [25]. ................ 24
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Figure 2.1 ACCC (Avionic Central Control Computer)............................................ 26
Figure 2.2 Dewetron measuring device. .................................................................... 27
Figure 2.3 Dewetron measuring device (continued). ................................................. 27
Figure 2.4 Fast response surface-mount thermocouple.............................................. 28
Figure 2.5 Thermocouples on the left wall. ............................................................... 29
Figure 2.6 Thermocouple in the cooling channel...................................................... 30
Figure 2.7 Detailed view of thermocouples on the right wall.................................... 31
Figure 2.8 All thermocouples on both sides. ............................................................. 32
Figure 2.9 VME cards in slots. .................................................................................. 32
Figure 2.10 Experiment setup. ................................................................................... 33
Figure 2.11 Experiment results of 25 ºC operating condition.................................... 35
Figure 2.12 Temperatures of slot5 and slot5_c while ACCC does not operate......... 36
Figure 2.13 Temperatures of slot5 and slot5_c while ACCC operates...................... 37
Figure 2.14 Temperatures of opposite walls. ............................................................. 38
Figure 2.15 ACS Test chamber.................................................................................. 40
Figure 2.16 55 ºC non-operating measurement results. ............................................. 41
Figure 2.17 55 ºC operating measurement results. .................................................... 42
Figure 3.1 Flow chart of DMAP. ............................................................................... 45
Figure 3.2 Chassis cooling method. ........................................................................... 46
Figure 3.3 General cooling method of convection..................................................... 47
Figure 3.4 Viscous sub layer...................................................................................... 49
Figure 3.5 Fan Curve Data Plot. ................................................................................ 51
Figure 3.6 Trend Line Added Fan Curve Plot (DMAP). ........................................... 52
Figure 3.7 Results of TMT for the operating conditions 25 ºC and 55 ºC................. 64
Figure 3.8 General view of DMAP in the chamber ................................................... 66
Figure 3.9 Walls numbers. ......................................................................................... 67
Figure 3.10 View factors............................................................................................ 68
Figure 3.11 Propimax 2 Fan....................................................................................... 73
Figure 3.12 Trend Line Added Fan Curve Plot (ACCC). .......................................... 74
Figure 4.1 3D Model of DMAP. ................................................................................ 78
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Figure 4.2 Computational Model of DMAP. ............................................................. 79
Figure 4.3 Meshed view of the chassis. ..................................................................... 81
Figure 4.4 Mesh Control Parameters. ........................................................................ 81
Figure 4.5 Detailed view of meshed fin. .................................................................... 82
Figure 4.6 Quality of the mesh................................................................................... 82
Figure 4.7 Advanced solver setup window. ............................................................... 83
Figure 4.8 Control points. .......................................................................................... 85
Figure 4.9 Wall temperature contour view of 25 ºC analyses (DMAP). ................... 86
Figure 4.10 Wall temperature contour view of 55 ºC analyses (DMAP). ................. 87
Figure 4.11 3D Model of ACCC................................................................................ 89
Figure 4.12 3D Model of ACCC (continued). ........................................................... 89
Figure 4.13 Meshed view of the chassis. ................................................................... 90
Figure 4.14 Wall temperature contour view of 25 ºC analyses (ACCC). .................. 92
Figure 4.15 Wall temperature contour view of 55 ºC analyses (ACCC). .................. 94
Figure A.1 Front section view of cooling channels. ................................................ 102
Figure A.2 Side view of cooling channel................................................................. 103
Figure B.1 Fan performance curve of Aximax 2…………………………………..104
Figure B.2 Dimensions of Aximax 2…………………………………...…………105
Figure B.3 Fan performance curve of Propimax 2……………………..………….106
Figure B.4 Dimensions of Propimax 2…………………………………………….107
xvii
LIST OF SYMBOLS
Latin Symbols
•
Q : Rate of heat transfer
airmm••
= : Cooling air mass flow rate
ieair TTT −=∆ : Temperature difference of cooling air exit and inlet
iT : Cooling air inlet temperature
eT : Cooling air exit temperature
pairp CC = : Specific heat capacity of air
n : Fin number
b : Fin thickness
a : Distance between two plate fins
H : Channel height
L : Channel length
w : Channel width
P : Wetted perimeter of cooling channel
allP : Total perimeter of cooling channel
allA : Total convection surface area
sA : Section flow area of one channel
sallA : All section flow area of one channel
HD : Hydraulic diameter
airV : Velocity of cooling air in the channel
xviii
VFR : Volumetric flow rate
Re : Reynolds number of channel cooling air
f : Friction factor of channel
Nu : Nusselt Number
h : Convection heat transfer coefficient
me : Efficiency equation coefficient
fA : Single channel surface area
lnT∆ : Logarithmic mean temperature difference
wT : Wall temperature
Pr : Prandtl Number
airk : Thermal conductivity of air
alk : Thermal conductivity of aluminum
calQ : Calculated heat dissipation rate
Greek Symbols
airρ : Air density
fη : Single fin efficiency
0η : All fins efficiency
airµ : Dynamic viscosity
1
CHAPTERS
CHAPTER 1
1 INTRODUCTION
In this thesis, a correlation based Thermal Model Tool (TMT) is developed for
standard Air Transport Rack (ATR) chassis and using the developed TMT, thermal
designs of standard ATR chassis are done in Aselsan Inc.
An ATR chassis is an avionic box. Design of a chassis starts with the given
geometrical and thermal constraints and mainly has three steps. At the first step, the
three dimensional (3D) model is generated. At the second step, the geometry is
numerically analyzed to determine if the given heat loads can be dissipated. Finally,
the designed model is produced and experimental studies are performed in order to
make the verification of analytical and numerical studies.
This thesis mainly consists of five chapters as; introduction, experimental studies,
analytical studies, numerical (Computational Fluid Dynamics (CFD)) studies and
discussion. In the introduction chapter, military ATR chassis are introduced. Types
of ATR chassis, dimensions, electronic card types, and cooling fans are explained. In
addition, new design ATR chassis with the name Digital Moving Map (DMAP) is
introduced. In the experimental studies chapter, results of the necessary experiments
that are done on the existing chassis are presented. In the analytical studies chapter,
by using the results of the experimental studies, mathematical model for the chassis
are prepared and TMT is developed. Cooling channel details of the DMAP are
studied by using TMT and necessary plate fin numbers are determined for different
operating conditions. Moreover, an existing chassis heat dissipation capacity is
2
calculated using TMT for a second case. In the CFD studies chapter, thermal
behavior of the created new design and existing chassis are analyzed with a CFD
tool. Finally, in the discussion chapter, results of TMT and CFD studies are
compared and future work subjects are stated.
VME cards are the compositions of Printed Circuit Boards (PCBs) and electronic
components. Example of an ATR chassis and a VME card are shown in Figure 1.1
and Figure 1.2 respectively. According to the standards, various chassis and VME
cards were manufactured in Aselsan Inc., previously.
Figure 1.1 Example of a standard ATR chassis [3].
Figure 1.2 Example of a standard VME card [3].
3
1.1 ATR Chassis
ATR chassis are military enclosures to provide a working environment to the
electrical/electronic equipment against the environmental factors. These types of
enclosures must have specific dimensions and interfaces, which are described in the
military standard called “ARINC 404A, Air Transport Equipment Cases and
Racking”. The reason of having a standard is to ensure the interface compatibility
between the manufacturers. It is expected that a device with a specific case
dimension produced by a manufacturer will fit in a mounting produced by another
manufacturer [1].
All interface types and envelope dimensions of the chassis are described in ARINC
404A [1]. Standard ATR case dimensions are given in Figure 1.3 and Table 1.1.
Figure 1.3 Standard ATR case dimensions [23].
4
Table 1.1 Standard ATR case dimensions [1].
Approx. Volume
Width(W) Length(L1) Length(L2) Height(H) ATR Size
[liter] [±0.76mm] [±1mm] [mm] [mm]
1/4 Short 3.52 57.15 318 320.5 193.5
1/4 Long 5.49 57.15 495.8 498.3 193.5
3/8 Short 5.57 90.41 318 320.5 193.5
3/8 Long 8.69 90.41 495.8 498.3 193.5
1/2 Short 7.7 123.95 318 320.5 193.5
1/2 Long 11.88 123.95 495.8 498.3 193.5
3/4 Short 11.8 190.5 318 320.5 193.5
3/4 Long 18.36 190.5 495.8 498.3 193.5
1 Short 15.96 257.05 318 320.5 193.5
1 Long 24.75 257.05 495.8 498.3 193.5
1 1/2 Long 37.62 390.65 495.8 498.3 193.5
Notes: Per ARINC characteristic 561 INS. the standard dimension ‘H’ = 193.5 mm may be increased to a maximum ‘H’ dimension of 269.88 mm when necessary for equipment reasons.
Military enclosure systems have difficult heat management problems and specific
environmental factors affect these enclosures. Therefore, cooling is essential for
these types of enclosures. The purpose of cooling system is to maintain the internal
components of electrical/electronic equipment at temperatures, which will achieve a
long and predictable service life [1].
There are two basic cooling methods that are described for ATR chassis: Flow-
Through and Natural Convection. If the heat load is very high, Liquid Cooling
method can also be used in some applications to cool the ATR chassis. However, it is
very difficult to control the behavior of the liquid in high altitudes and accelerations,
so liquid cooling is not a preferred method for most of the cases. Therefore, Flow-
Through and Natural Convection are the most frequently used methods.
5
1.1.1 Flow-Through Cooling Method
The standard mode of cooling will be to draw air through the unit by applying a
differential pressure between the inlet and outlet sections of the unit. This normally
should consist of suction applied to the inlet of the unit to create air movement. This
type of cooling method has two subgroups:
• Direct forced convection cooling inside the chassis.
• Forced convection cooling via cooling channels.
In the first subgroup, the cooling air is forced inside the chassis, and the electronic
components are cooled directly by the cooling air. However, this type of cooling is
not applied if the ATR chassis is not positioned in the cockpit since the ATR chassis
must be insulated against environmental conditions. Otherwise, environmental
conditions will increase maintenance and reduce service life. Besides that, some
electronic components do not operate when exposed to moisture and this is another
reason for sealing.
In Figure 1.4, an example chassis is shown for the first subgroup of flow-through
cooling. Cooling air is sucked from the front panel of the chassis and flows through
the electronic cards. In this manner, generated heat is dissipated to the cooling air
directly. Finally, cooling air leaves the chassis with increased temperature.
Additionally, this chassis has mass and dimension advantages since there is no need
for cooling channels,
6
Figure 1.4 Flow-through cooling chassis [2]
In the second subgroup, cooling air does not flow through the electronic cards.
Cooling channels are created with fins and cooling air flows through these channels.
The generated heat is conducted to the walls of the cooling channel and the heat is
dissipated to the cooling air by convection. In this manner, the chassis is cooled
unaffected by the environment. Ability to place the chassis at any point in the air
vehicle is an advantage of this method.
However, existence external cooling channels increase both the dimensions and the
mass, which is a disadvantage of this method. In Figure 1.5, an example chassis with
cooling channels is shown.
1.1.2 Natural Convection Cooling Method
Heat dissipation of natural convection is lower than forced convection. In addition,
this type of cooling allows sealing. If the rate of generated heat is low and insulation
is required for the chassis, this cooling type should be preferred.
7
Figure 1.5 Forced convection cooling chassis with cooling channels [3]
In Figure 1.6, an example for the natural convection chassis is shown. In such a kind
of chassis, the generated heat is conducted to the chassis and the chassis is cooled by
natural convection.
Figure 1.6 Natural Convection cooling chassis [4].
8
The ATR chassis, which is the subject of this thesis, must have high heat dissipation
rates because of generated high heat values and the chassis must be sealed due to the
location in the helicopter. In addition, placement of the chassis is in the tail section of
the helicopter and may be affected by rain or other environmental conditions. These
two limitations identify the cooling method for the chassis. The chassis must be
sealed and must have high heat dissipation rates. Therefore, forced convection
method with cooling channels is convenient for the chassis. For this purpose, plate
fins or pin fins can be used inside the cooling channels. Plate fins are preferable to
pin fins for the advantages of low production cost, more convection surface area, and
directed flow. Furthermore, machining or casting plate fins are easier than pin fins.
There are side-cooling channels on the chassis. The draft model of the chassis with
side plate fins is shown in Figure 1.7.
Figure 1.7 Draft model of the chassis.
Ametek Rotron fans are used in Aselsan Inc. because they meet the military
requirements. One of the existing fans is selected also for this study. Fan is placed at
the rear side of the chassis as shown in Figure 1.8. Cooling air is sucked from suction
Cooling air inlet section (On both sides).
9
opening of the channels shown in Figure 1.7. Cooling air accumulates at the front
side of the fan and exhausts to the environment with high temperature.
Figure 1.8 Placement of Fan.
1.2 VME Cards
Printed Circuit Board (PCB) has conductive pathways and connects electronic
components electrically. Furthermore, PCB is used to mechanically support the
electronic components and is also called as Printed Wiring Board (PWB) or etched
wiring board. If a PCB is populated with electronic components, it is called as
Printed Circuit Assembly (PCA), also known as a Printed Circuit Board Assembly
(PCBA). Much more layout effort and higher initial cost are required for PCBs than
either wire wrap or point-to-point constructions. Besides this for high-volume
production, PCBs are much cheaper and faster to produce.
There are three basic VME card sizes, which are 9U, 6U, and 3U. 6U card has two
different models. All sizes are shown in Figure 1.9.
10
Figure 1.9 Standard VME card sizes [5].
The cooling of VME cards is done with thermal plates. Generated heat is conducted
to thermal plates and through them to the chassis. In Aselsan Inc., also thermal plates
are designed for the used VME cards. One of the designed thermal plates is shown in
Figure 1.10.
Figure 1.10 Thermal Plate.
11
There are four different VME cards in the new design chassis DMAP. One of them is
power card and the rest 3 are the processor cards each having a special mission. For
instance, the card with the special name “AIK” has the mission of image processing.
In Table 1.2, the list of the used VME cards is given with their power loads. In
addition, the placements of slots are shown in Figure 1.11. There is approximately
100W (28V x 3.5A) power load inside the chassis, which is generated by the VME
cards.
Table 1.2 Power loads of the VME cards.
Slot Number Card Name Power
Consumption [A]
Slot-1 AIK+1553+429 1.1
Slot-2 VIK 0.6
Slot-3 AIK 0.8
Slot-4 Power Card 1
Total heat load and geometrical details are parameters for the design. Cooling
channel dimensions are shown clearly in Appendix A. The height, width, and length
of the channel are 158 mm, 10 mm and 330 mm respectively. These are the available
dimensions for plate fin placements. By using the given geometrical constraints,
DMAP is modeled by using proEngineer tool as shown in Figure 1.8. Unknowns are
the plate fin details for the cooling channels. Analytical calculations are done in
order to find the sufficient number of plate fins for the given heat load by considering
given parameters for the cooling channels and ambient operating conditions 25 ºC
and 55 ºC.
12
Figure 1.11 Placement of slots.
Slot-1
Slot-2
Slot-3
Slot-4
13
1.3 Vaneaxial Fans
Two Ametek Rotron Aximax 2 (Figure 1.12) fan is used in the back of the chassis,
which is a Vaneaxial fan. Small Vaneaxial is a high-speed and compact fan type,
whose airflow is parallel to the motor shaft. For minimum acoustical noise and
maximum aerodynamic efficiency, the guide vanes and impeller are of airfoil
constructions. These fans are designed for use in airborne avionic boxes where sizes,
weight, and reliability are critical, and where high heat loads must be dissipated with
cooling air. Optional internal Fan Performance Sensor (FPS) or an external Low
Speed Warning Device (LSWD) is available for most units [12].
Figure 1.12 Aximax 2 fan [12].
Vaneaxial Fans are axial flow air moving devices. In these fans, the motor rotor is
cast inside the impeller to achieve the smallest possible axial dimension. Through
those true airfoil blade designs, the higher aerodynamic performance and efficiencies
can be achieved [12].
Those fans are military, compact, and quiet designs, which are used mostly for
cooling electronic enclosures under severe environmental conditions in airborne,
ground based and shipboard applications where air can move freely with low static
pressure. Axiamax 2 fan can be designed to meet the military standard MIL-STD-
14
810C requirements [12]. The general specifications of Aximax 2 are given in Table
1.3 and dimensions and critical performance curve are given in Appendix B.
Table 1.3 General Specifications of Aximax 2 [12].
Physical envelope: 55.3 mm Dia. x 42.2 Length
Weight: approximately 120 gr.
Die cast aluminum Venturi and Rotorprop
Specially designed for cooling electronics in aircraft, ground-based and
shipboard applications
All aluminum components finished with chemical conversion coating
per MIL-C-5541
Top coat of lusterless black enamel, color #37038, per Federal Standard
595 conforming to TT-E-489 Type B.
Corrosion-resistant stainless steel shaft and hardware.
Meets or exceed the requirements of MIL-B-23071 and other applicable
U.S. military and commercial aerospace specifications.
Max free delivery airflow of 59 CFM.
Ambient temperature range: -54 °C to 100 °C.
Acoustic levels as low as 55 dBA.
1.4 Literature Review
In the literature, there are many studies done on the plate fins related with the air-
cooled heat sinks. However, the number of studies done related with the heat sinks or
cooling channels placed on ATR chassis are not so much. In this section, studies
conducted by van Engelenhoven et al [9], Iyengar and Bar-Cohen [11], Wu et al [20],
Leon et al [24] and Fowler et al [25] are examined.
15
Jesse and Gary [9] had done one of the studies about ATR chassis. In the study, by
considering pressure drop requirement and fin manufacturing capabilities for
ruggedized military electronics, they examined chassis-level air-cooling limits. In
order to maximize the heat dissipation rate of an air-cooled chassis, they optimized
longitudinal plate fin (included in side wall ducts) geometry. For the optimization,
numerical and analytical models were developed. The results of the studies were
presented in the form of a performance map. Results showed the differences of
specified set of mass flow, pressure drop, and heat transfer requirements between
different fin manufacturing processes. According to the results, if isothermal
boundaries could be achieved instead of isoflux boundary condition assumption, the
heat transfer capacity of the chassis would increase. Studies were done on different
manufacturing process given in Table 1.4.
One of the results of the study shows the influence of fin thickness on thermal
resistance (Figure 1.13). Studies were done for different edge temperatures as 55 ºC,
70 ºC and 85 ºC and for two different chassis types as long and short. One of the
performance maps is given in Figure 1.14. To maximize heat transfer in all cases
bonded or folded fins with a thickness of 0.254 mm were obtained.
Table 1.4 Fin manufacturing limits based on Iyengar [10].
Process Min & [mm]
Extrusion 1.575
Forging or Gang saw 1
Skiving 0.6
Swaging 0.5
Bonded 0.254
Folded 0.05
16
Figure 1.13 Fin thickness and number influence on fin stack thermal resistance [9]
Figure 1.14 Short chassis, top to bottom flow, 250Pa pressure drop, 85C edge [9].
17
Iyengar and Bar-Cohen [11] made the next study. In order to find the maximum heat
transfer capabilities of the heat sinks, analytical model, and the thermofluid
performance of the heat sink was characterized. In addition, least-material
optimization was done to achieve optimal material usage. Different production
methodologies (extruded, die-casting, bonding, folding, modified die-casting, skiving
and machining) were examined. The manufactures with different methods are given
in Table 1.5 and parameters were defined in Figure 1.15.
Figure 1.15 Side-inlet-side-exit (SISE) rectangular plate fin heat sink conf. [11].
Table 1.5. Manufacturability Constraints [11]
In the study, to identify the maximum heat transfer capability, optimization
procedure was adopted. A least-material optimization was done. With respect to
18
mass-specific heat dissipation, magnesium was found to be the most efficient
material.
The next examined study had done by Wu et al. [20] predicts the hydraulic and
thermal performance of a plate-fin heat sink. They developed all-in-one asymptotic
model for a wide range of Reynolds numbers, including laminar, transition, and
turbulent flows as Re<5000. The model can predict pressure drops with accuracy
within -13.87% to 8.4%. They developed a practical model for convectional plate-fin
heat sink. The model derives a pressure drop correlation for the working fluid.
Figure 1.16 Practical model of a conventional plate-fin heat sink [20].
They divided the pressure drop inside the heat sink into two parts as the friction term
and the term due to the change of flow section. The heat sink pressure drop was
considered as in Eqn. 1.1, where fapp is the fanning friction factor, Kc and Ke are the
contraction and expansion pressure loss coefficients. V is the velocity of coolant; L
and Dh are the channel length and hydraulic diameter, respectively.
2
2
14 VKK
D
LfP ec
h
app ρ
++=∆ ( 1.1)
19
They generated an asymptotic correlation for the general friction factor as in the Eqn.
1.2.
( ) nn
turb
n
lamapp fff/1
+= ( 1.2)
Because all the flows in the thesis study are turbulent, only the ( ) nn
turbf/1
part of the
Eqn. 1.2 is used. They gave the equation of this part as in Eqn. 1.3.
175.02.0Re0962.0
−
−⋅=h
c
appD
Lf ( 1.3)
Also they gave the equations for Kc and Ke as Eqns. 1.4 and 1.5.
2
4.08.0
−=
p
aK c ( 1.4)
−
−=
p
a
p
aK c 4.01
2
( 1.5)
Leon et al. [24] worked on the influence of cooling channel fin shape on the pressure
drop caused by flow resistance of a heat sink. They developed a new manner to
optimize the heat sink. They emphasized not only on maximum heat transfer flux,
but also on minimum flow resistance with the developed manner. They studied
numerically using the computational fluid dynamic software FLUENT.
They found the advantages of using aerodynamic shaped fins if the Reynolds
number, is greater or equal than about 800. The authors suggested some preliminary
profile shapes and they considered this research as a unique approach if the study is
compared to other studies where the flow resistance has not been taken into account.
20
The authors studied on 4 different fin geometries. Figure 1.17 shows a standard
arrangement of rectangular fins. Figure 1.18 shows fins with an aerodynamic shape
(airfoil shape). The purpose of this arrangement was stated as to lower the
aerodynamic drag. The cooling fin area was the same as in the rectangular
arrangement. Figure 1.19, and Figure 1.20 show fins where the inlet edge or the inlet
and outlet edges are rounded.
Figure 1.17 Rectangular model: T0= 300 K, Tw =423 K [24].
Figure 1.18 Aerodynamic model: T0= 300 K, Tw =423 K [24].
Figure 1.19 Rounded inlet model (Rounded I): T0= 300 K, Tw =423 K. [24].
21
Figure 1.20 Rounded inlet and outlet model: T0= 300 K, Tw =423 K [24].
As conclusion, the authors stated that, if the Reynolds number was greater or equal
than about 800, the reduction of the flow resistance could be achieved by the use of
aerodynamic profiles. In this manner, the value of the removed heat was not affected.
They suggested that as a practical optimum, a cooling fin with a rounded leading
edge.
They concluded that, for very small Reynolds numbers around 100 or less, they
found that it was useless to introduce any type of aerodynamic layout. Flow resistant
reduction would permit to use of lighter fans without affecting the removed heat. As
a result, noise level of the fan, fan size and power consumption would be reduced.
Lastly, they stated that the given conclusions are valid for any type of heat sink
where the flow is parallel to the fins.
Fowler et al. [25] studied the optimal geometric arrangement of staggered parallel
plates in a fixed volume with forced convection heat transfer experimentally and
numerically. They tried to maximize the total heat transfer rate when the maximum
temperature at a point inside the volume cannot exceed a certain level.
The study was done in two parts as experimental and numerical. Experimental results
are reported for air in the range 1000 < Re < 6000, where L is the swept length of the
fixed volume. They supported the findings with numerical results to 100<Re< 10000.
In addition, they showed that there is an optimal way to stagger the plates. The
geometric parameters are shown in Figure 1.21.
22
Figure 1.21 The geometric parameters [25].
Authors investigated the effect of changing the stagger parameter β in two ways as in
Figure 1.22 and Figure 1.23.
They interpreted Figure 1.22 as “the spacing W was set equal to the optimal value
Wopt that corresponds to perfectly staggered plates (β=1). As the stagger parameter
varies over [0, 1] interval, the overall thermal conductance exhibits a maximum at a
certain β value (called βopt). This maximum becomes sharper as A decreases below
A=1: in this range the stagger associated with maximum q is closely approximated
by βopt=A. Perfectly staggered plates (β=1) are always better than in-line plates
(β=0); however, β=1 is not the optimal stagger parameter when A<1. Figure 1.22
also shows that when A>1 the effect of β on q is relatively insignificant” [25].
23
Figure 1.22 The effect of the stagger when the spacing is fixed at Wopt (β = 1) (N=4,ReL=103) [25].
They interpreted Figure 1.23 as the alternative to determine the optimal spacing that
corresponds to each value of the stagger parameter, namely Wopt(β). This route was
followed in the construction of Figure 1.23, where A = 0.5, N = 4 and ReL = 103. The
optimal spacing increases by roughly 7% as β decreases from 1 to 0. In other words,
the optimal spacing for in-line plates is slightly larger than for perfectly staggered
plates. The corresponding β effect on the maximized overall conductance is also
small (within 6%). There is a weak q maximum at β = 0.5, and, as might be expected,
we learn again that perfectly staggered plates (β = 1) perform better than in-line
plates (β = 0)” [25].
24
Figure 1.23 The effect of the stagger parameter on the optimal spacing and the maximum overall
thermal conductance (A = 0.5, N = 4, ReL = 103) [25].
Jouhara and Axcell [26] studied the thermal conditions within a heat sink with
rectangular fins under laminar forced convection cooling. To model the increase in
air temperature through channels, classical heat transfer theory and a computational
approach were used. They represented the variation of the key heat transfer
parameters with axial distance, the rapid changes in heat transfer coefficient fin
efficiency near the leading edges of the cooling fins. They described the
mathematical modeling and solution techniques for each method in detail.
Four different approaches were used during their study:
• Idealized case: In this case study they assumed that the fin surfaces were at a
uniform temperature equal to that of the base plate.
• Approximate analysis: Heat transfer with non-uniform wall temperatures.
• Analysis using numerical integration along the flow passage: They assumed
that both the heat transfer coefficient and the fin efficiency vary along the
flow.
25
• Finite element/CFD study: The computational study was performed.
Authors obtained the results for each of the different theory. They presented the
predicted thermal performance results of a single heat sink using the four methods,
which is given in Table 1.6 for approach velocities from 1 to 8 m/s.
Table 1.6 Thermal performance of heat sink [26].
Initially, they presented for idealized fins which are 100% efficient using an
expression for mean Nusselt number. Next, they utilized a uniform fin efficiency
based on the mean Nusselt number. Then they introduced more detail by allowing
both Nusselt number and fin efficiency to vary along the flow passage. Finally, they
modeled the effects of flow disturbance at entry to the heat sink and axial conduction
in the cooling fins in a CFD.
They concluded the study as stating that “although the calculations show that heat
transfer coefficients and fin efficiencies vary substantially along the flow path, good
engineering accuracy for heat sink performance can be obtained for laminar flow
using calculations in which mean values for the Nusselt number and fin efficiency
are calculated from analytical expressions” [26].
26
CHAPTER 2
2 EXPERIMENTAL STUDIES
Aim of the experimental studies is to obtain critical wall temperature values for
different operating conditions. The critical wall temperatures are used in analytical
calculations to develop TMT and to obtain appropriate number of fins for the chassis
DMAP.
There are existing chassis in Aselsan Inc. used for different applications.
Experimental studies are conducted on one of the existing chassis, which is an
Avionic Central Control Computer (ACCC) of a helicopter, shown in Figure 2.1.
Figure 2.1 ACCC (Avionic Central Control Computer).
27
In the experimental studies, temperatures of walls and cooling air are measured by
Dewetron data acquisition system shown in Figure 2.3. The system has 32 channels
and the sampling rate of the system is set to be 1s. Moreover, two thermocouple
interface pads are utilized to connect thermocouples to the system.
Figure 2.2 Dewetron measuring device.
Figure 2.3 Dewetron measuring device (continued).
28
Fast response surface-mount thermocouples (Figure 2.4) are used for the
measurements of wall temperatures. The thermocouple has a bare sensing element
with a nominal 0.025 mm thick, 25.4 mm length and 9.5 mm width [18]. 30 AWG
(0.0254 mm) thermocouple wire with a temperature sensitivity of ± 1.1 °C is used for
the fast response thermocouples [19].
Figure 2.4 Fast response surface-mount thermocouple.
Since generated heats are different in the slots, wall temperatures of the slots are
different as well. However, in the analytical calculations, a single wall temperature is
used. Because of that the mean temperature of all slots are necessary. Therefore, in
order to get temperature degrees of each slot, thermocouples are used on slots as
shown in Figure 2.5. Due to the empty slots 8 and 9, thermocouples are not placed on
these slots. Temperature values of each slot are measured on the left wall. Average
temperature of slots is calculated to obtain critical wall temperature.
29
Figure 2.5 Thermocouples on the left wall.
2.1 Assumptions
Before starting experiments, some assumptions are done in order to facilitate the
studies. The assumptions are listed as below:
• Inside and outside temperatures of the plate finned wall are close to each
other.
• Left and right side wall temperatures are close to each other and symmetry
condition is acceptable.
2.1.1 Assumption-1
Firstly, assumed that inside and outside temperatures of the plate finned wall is close
to each other. Although the convection occurs between outside wall of chassis and
Slot 11
Slot 10
Slot 7
Slot 6
Slot 5
Slot 4
Slot 3
Slot 2
Slot 1
30
the cooling air, it is not easy to make proper temperature measurements at the outside
wall, as fan is operating. In such a situation, cooling air also flows over the
thermocouples and cools them. This causes to measurement errors. Therefore,
assumed that inside and outside temperatures of wall are close to each other because
of thin (2 mm) wall thickness. In order to verify this assumption, an additional
thermocouple is placed on the cooling channel as shown in Figure 2.6. This
thermocouple coincides with 5th slot thermocouple, which is inside the chassis. The
temperature measurements of both thermocouples are compared to see if the
assumption is true. Detailed verification of the assumption is given in Section 2.2 and
Figure 2.12.
Figure 2.6 Thermocouple in the cooling channel.
2.1.2 Assumption-2
Second assumption is the symmetry condition. Assumed that left and right cooling
channels have the same cooling capacity and symmetry condition is applicable. By
This thermocouple coincides with 5th slot thermocouple.
31
considering the fragile features of the thermocouples and placements difficulties,
measurement are done for one of the sidewalls. Therefore, to verify this assumption
and understand general characteristics of the sidewalls, additional thermocouples are
placed on the right wall of the ACCC (Figure 2.7).
Figure 2.7 Detailed view of thermocouples on the right wall.
Three thermocouples are stuck on 1st, 5th and 11th slots of the right wall. The
temperature values measured with these thermocouples are compared with same slot
left wall thermocouple measurements. For instance, left wall 1st slot thermocouple
measurement is compared with right wall 1st slot thermocouple measurement. The
thermocouples on the opposite side of the slots are named as Slot 1o, Slot 5o and Slot
11o as shown in Figure 2.7. Results are examined and temperature differences of left
and right wall are determined. All placed thermocouples are shown in Figure 2.8.
Slot 1o Slot 5o Slot 11o
32
Figure 2.8 All thermocouples on both sides.
There are 3 additional thermocouples in the battery case of the ACCC. Temperature
limit of in use battery is 100 °C maximum, so with these thermocouples, battery case
temperature is identified if there is any challenge. After thermocouples are stuck on
the related slots, electronic card of each slots are placed as shown in Figure 2.9.
Figure 2.9 VME cards in slots.
Battery Case
Left Wall
Right Wall
33
19 thermocouples are utilized during experiments at below given positions:
• 9 pieces, on the left wall
• 3 pieces, on the right wall
• 3 pieces, in the battery case
• 1 piece on the left wall air side
• 1 piece for operating condition
Cooling air inlet temperature is measured by the thermocouple allocated for
operating condition. Thermocouples are put out from the chassis using “Elapsed
Time Meter (ETM)” and “Grounding Screw” holes as shown in Figure 2.10. The
general view of experiment setup is also shown in Figure 2.10.
Figure 2.10 Experiment setup.
Grounding Screw Hole
ETM Hole
34
The experiments are built for two main conditions. Firstly, 25 °C ambient cooling air
experiment is performed. In this experiment, inside and outside temperatures
differences of the wall are examined. It is important to indicate that fan does not
operate in this measurement. Hereby it is showed that measuring wall temperatures
from inside of the wall is convenient. In the next step, left and right wall
temperatures are evaluated to show that both wall temperatures are close to each
other, so analytical calculations can be done by using temperature measurements of
one wall with negligible errors. Secondly, the same experiment is performed at 55 °C
ambient temperature and the same measurements are done for this alternative as well.
2.2 25 °C Ambient Temperature Experiments
This experiment is performed at 25 °C ambient air temperature. Experiment duration
is approximately 110 minutes. 19 thermocouples are used for data collection.
Approximately 60 minutes later, ACCC reaches to steady state condition and
measured temperatures do not change any more. Power input to the ACCC is
calculated by using power supply current rate input. This power load can be thought
as the heat load on the system, which must be dissipated. Up to 25th minute, power
supply shows that 6.1 A current rate with 28 V. This shows that power input to
ACCC is approximately 170 W. All cards placed in the chassis with the order stated
in Table 2.1.
Experiment results are analyzed and temperature values are plotted as shown in
Figure 2.11. ACCC is cooled by 2 fans and these fans are controlled with a fan
control card. This card is placed at the bottom of the ACCC and has a relay on it. Fan
control card starts up the fans when the relay temperature reaches to 45 °C.
Therefore, fans must be operating later than ACCC.
35
Table 2.1 Power loads of cards
Slot Number
Card Name Power
Consumption [A]
Slot-11 Power Card 1.1
Slot-12 Power Card 0.0
Slot-1 AIK+1553+429 1.2
Slot-2 AIK+GIK+GIK 1.1
Slot-3 VIK 0.7
Slot-4 AIK+GIK+GIK 1.0
Slot-5 AIK 0.8
Slot-6 AAK 0.1
Slot-7 AAK 0.1
Slot-10 GDK 0.0
Experiment results show that fans start to operate 25 minutes later than ACCC. Up to
minutes 25, fans are turned off by fan control card and all temperatures increase.
When the relay of fan control card reaches to 45 °C, fans start to operate and all
temperatures start to decrease.
Slot Temperatures - Time
25
26
27
28
29
30
31
32
33
34
35
36
37
38
39
40
41
42
43
44
45
0 10 20 30 40 50 60 70 80 90 100 110
Time [Min.]
Tem
pera
ture
[°C
]
slot1 slot2 slot3 slot4 slot5
slot6 slot7 slot10 slot11
Figure 2.11 Experiment results of 25 ºC operating condition.
36
In order to see the inside and outside temperature differences of left wall, cooling
data of ACCC are also collected. In this experiment, ACCC is not operating and left
for cooling. Time scaled cooling temperatures graph is given in Figure 2.12. Results
show that slot5 temperature is closed to slot5_channel temperature. This
measurement shows that, measuring wall temperatures from inside is reasonable.
Wall Temperatures - Time
27
28
29
30
31
32
33
34
35
36
37
0 5 10 15 20 25 30 35 40 45 50 55 60
Time [Min.]
Tem
pera
ture
[°C
]
slot5 slot5_c
Figure 2.12 Temperatures of slot5 and slot5_c while ACCC does not operate.
The reason of doing this measurement while ACCC is out of service is not to affect
thermocouples by flowing air. There is also an additional measurement data taken
while experiment started. This data is shown in Figure 2.13. Critical minute 25 is
marked. Up to this point, fans do not operate and again slot5 and slot5_c
thermocouple values are closed to each other. However, when fans started to operate
it is shown that there are differences between measured temperature values. As
37
stated, the reason of the difference can be thought as the affect of cooling air on
slot5_c thermocouple.
Therefore, first assumption is verified and measuring wall temperatures from inside
of the wall does not cause to inaccuracy. Quite the contrary, wall inside measurement
gives more reliable results than wall outside measurement.
Wall Temperatures - Time
24
26
28
30
32
34
36
38
40
42
0 10 20 30 40 50 60 70 80 90 100 110 120
Time [Min.]
Tem
pera
ture
[°C
]
slot5 slot5_c
25
Figure 2.13 Temperatures of slot5 and slot5_c while ACCC operates.
Also second assumption must be verified if it is possible to continue experiments by
measuring only left wall temperature values. Because in analytical calculations,
dissipated heat is calculated for only one wall by assuming symmetry condition is
applicable.
38
To examine if symmetry condition is convenient right wall slot temperatures are
compared with left wall slot temperatures. Right wall 1st, 5th and 11th slot
temperatures are compared with the left wall 1st, 5th and 11th slot temperatures
respectively. In Figure 2.14 results of these measurements are given.
Wall Temperatures - Time
25
27
29
31
33
35
37
39
41
43
0 10 20 30 40 50 60 70 80 90 100 110
Time [Min.]
Tem
pera
ture
[°C
]
slot1 slot5 slot11
slot1o slot5o slot11o
Figure 2.14 Temperatures of opposite walls.
Summary of above graphs is tabulated as below. Temperature differences between
slot1, slot5 and slot11 thermocouples are 2.1 °C, 2.2 °C and 0.8 °C respectively.
However, measured data show that left wall of the ACCC is hotter than right wall;
temperature differences do not constitute big errors in analytical calculations. On the
other hand, analytical calculations will be done for the more critical left wall
temperature values.
ACCC reaches to steady state condition approximately 60 minutes later as shown in
Figure 2.13. The steady state slot temperatures are tabulated Table 2.3. There are
39
totally 9 thermocouples on the slots and the average value of measured temperatures
is calculated as 33.3 °C for the left wall.
Table 2.2 Opposite channels thermocouple results.
Thermocouples Temperatures [°C] ∆T [°C]
slot1 (L) 33.7
slot1o (R) 31.6 2.1
slot5 (L) 33.8
slot5o (R) 31.6 2.2
slot11 (L) 30.7
slot11o (R) 29.8 0.8
As it is seen both in Figure 2.14 and Table 2.2, slot temperatures are distinct from
each other as a result of different power loads of electronic cards (details of heat
loads of each card is given in Table 2.1). These various power loads are converted to
the heat loads, which cause different temperature levels on the slots.
Table 2.3 Measured slot temperatures at 25 ºC.
Slot1 Slot2 Slot3 Slot4 Slot5 Slot6 Slot7 Slot10 Slot11 Average
Steady State Temperatures [°C]
33.7 36.6 34.7 34.6 33.8 32.7 31.8 30.8 30.7 33.3
2.3 55 °C Ambient Temperature Experiments
This experiment is conducted at 55 °C ambient air temperature in a test chamber,
which is shown in Figure 2.15. ACCC is placed in the chamber and the temperature
of the chamber is set to 55 °C.
40
The necessary condition in order to conduct the experiment is convergence of ACCC
to 55 °C. Data collection is started and measurements are saved in order to confirm if
ACCC reached to 55 °C steady state condition. As it is seen in Figure 2.16, all
thermocouples are converged to 55 °C approximately 155 minutes later. Actually,
155 minutes time window is the convergence of ACCC to 55 °C and does not
operate during the convergence time window.
Figure 2.15 ACS Test chamber.
41
Temperature - Time
20
25
30
35
40
45
50
55
60
0 20 40 60 80 100 120 140 160
Time [Min.]
Tem
pera
ture
[°C
]
slot1 slot2 slot3 slot4 slot5 slot6 slot7 slot10 slot11
Figure 2.16 55 ºC non-operating measurement results.
Experiment duration is approximately 75 minutes. 19 thermocouples are used for
data collection. Approximately 60 minutes later ACCC reaches to steady state
condition and measured temperatures do not change any more.
According to experimental results, steady state slot temperatures are tabulated in
Table 2.4. Most critical temperature is shown in the Slot2 which is 66.7 °C. Average
temperature of the slots is calculated as 63.3 °C.
42
Table 2.4 Measured slot temperatures at 55 ºC.
Slot1 Slot2 Slot3 Slot4 Slot5 Slot6 Slot7 Slot10 Slot11 Average
Steady State Temperatures [°C]
63.6 66.7 64.7 64.7 63.7 62.6 61.8 60.6 61.0 63.3
Besides all these measurements, critical CPU junction temperature values are also
collected. It is seen that for this chassis, CPUs work on the critical temperature limit,
which is 95 ºC. Therefore, measured wall temperature values are the design
considerations for the next studies.
Temperature - Time
50
51
52
53
54
55
56
57
58
59
60
61
62
63
64
65
66
67
68
0 5 10 15 20 25 30 35 40 45 50 55 60 65 70 75 80
Time [Min.]
Temperature [°C]
slot1 slot2 slot3 slot4 slot5 slot6 slot7 slot10 slot11
Figure 2.17 55 ºC operating measurement results.
43
CHAPTER 3
3 ANALYTICAL STUDIES
In Aselsan Inc., in particular, but also in the most of the electronic industry, in
generally, design of an ATR chassis is done with the cooperation of Mechanical
Design and CFD Analysis departments. Mechanical design engineers decide about
the cooling channel details by using the previous design experiences. For instance, a
mechanical design engineer tries to apply the plate fin details of previous chassis to
the new design without any knowledge if the same plate fin details are applicable to
the new conditions. After design is completed, the model is transferred to CFD
analysis team to understand if the cooling capacity of the chassis is sufficient or not.
Generally, new designs do not meet necessary cooling requirements. At this stage,
analysis team gives some feedback to the mechanical designer in order to change the
cooling channel details. This change is done by increasing or decreasing the number
of plate fins, by changing the width of the channel or by changing the thickness of
the plate fins. All these parameters are unknowns and there are many options with
different combinations. At this stage, there is an uncertainty and the mechanical
design engineer does not know which way to go. The CFD analysis team must verify
every new option tried by the mechanical design engineer before production and
every loop between the design engineer and the analysis team takes several weeks.
This trial and error procedure between the design engineer and the analysis team
causes time and money loses.
44
The aim of the analytical studies is to develop a Thermal Model Tool (TMT), which
will be used by mechanical design engineers before starting to the mechanical
design. By using this tool, initial thermal design process of the ATR chassis will be
completed in order to make more affective designs.
During analytical studies, thermal mathematical model of the chassis is generated.
The generated thermal model can be carried out not only for DMAP but also for any
type of ATR chassis. The input of the model is critical wall temperatures obtained
from experimental studies and geometrical limitations. By using given limitations,
the mathematical model will give a basic feedback to mechanical designer about the
cooling channel geometry of the chassis. The mechanical designer uses the feedback
and checks if it is possible to dissipate the generated heat within the chassis, which
will be formed with given constraints.
In the analytical calculations chapter, two different ATR chassis are considered. One
of them is the new design chassis DMAP and the second one is the existing chassis,
on which the experimental studies were carried out, ACCC. Thermal model tool is
developed on DMAP and analytical calculations are done for both DMAP and
ACCC. Although ACCC is not a new design chassis, the aim of the analytical studies
for ACCC is to test the TMT on a different chassis.
3.1 DMAP Analytical Studies
As mentioned in Page 8, forced convection method is used, and the chassis has
cooling channels. General flow chart of the chassis is shown in the given cross
sectional view Figure 3.1. Cooling air is sucked from the side air intake openings,
flows via the plate-fins, collected in front of the fan, and exhausted to the
environment.
45
Figure 3.1 Flow chart of DMAP.
Cooling air opening
Cooling air opening
Plate Fins Plate Fins
Cooling air coolectin section
46
Figure 3.2 shows the cooling method of chassis. The generated heat is conducted
from central processing units (CPU) to thermal plate and then conducted to slot of
chassis. Similarly, heat is conducted to finned structure. Cooling air is flowing
through finned structure; in this way heat is dissipated by convection to the cooling
air. Because the same VME size electronic cards are used, sizes of the thermal plates
are also the same for all of the cards. Moreover, behavior of the thermal plates is
similar and known from previous studies. For this reason, convection section of the
cooling is considered in mathematical model and conduction is not studied.
Figure 3.2 Chassis cooling method.
On the electronic cards, CPU is the most critical component and this study is based
on the critical operating temperature of the CPUs. Critical junction temperature for
the CPUs is 95 °C. By using the results of experiments, it is learnt that the average
slot wall temperature is about 63.3 ºC when the CPU reaches to critical junction
temperature (95 °C). This case occurs at 55 °C operating condition. In the
mathematical model of the chassis, well-known empirical connection correlations are
PCB, E. component, thermal plate
Cold Air
Air Channel
Slot
Fin Structure
47
used. At the end of the analytical studies, radiative heat transfer is examined also to
understand if the rate of transferred heat by radiation is negligible or not.
Analytical calculations are observed under two topics; heat transfer calculations and
pressure drop calculations.
3.1.1 Heat Transfer Calculations
General cooling method of convection is summarized in Figure 3.3. Heat transfer to a
fluid flowing in a tube is equal to the increase in the enthalpy of the fluid. Energy
balance equation can be written as:
)( iep TTCmQ −=••
( 3.1)
This is the cooling capacity of flowing air inside the cooling channel and equal to the
dissipated heat load. Dissipated heat is transferred to cooling air by convection. So
energy balance equation can be written as:
lnThATCmQ channelairairpair ∆=∆=••
( 3.2)
Figure 3.3 General cooling method of convection
48
•
Q = heat load to be dissipated
airmm••
= = cooling air mass flow rate
ieair TTT −=∆ = temperature difference of cooling air exit and inlet
iT = cooling air inlet temperature
eT = cooling air exit temperature
airpp CC = = specific heat capacity of air
h = convection heat transfer coefficient
channelA = surface area of convection
lnT∆ = logarithmic temperature difference
In the mathematical model, by using heat load (generated heat), necessary cooling
surface area, or in other words necessary number of plate fins are calculated. In
calculations, most critical variable is convection heat transfer coefficient and can be
calculated from Nusselt number.
4.0=n for heating and 3.0=n for cooling. nNu PrRe023.0 8.0= ( 3.3)
This equation is known as Dittus-Boelter equation [6].
Eq. 2.3 is fairly simple, and it may give errors as large as 25 percent. This error can
be reduced considerably by using more complex but more accurate relations such as
the second Petukhov correlation [6] expressed as:
)1(Pr)8/(7.1207.1
PrRe)8/(3/25.0 −+
=f
fNu ( 3.4)
49
The accuracy of this relation at lower Reynolds numbers is improved by modifying it
as [6]:
)1(Pr)8/(7.121
Pr)1000)(Re8/(3/25.0 −+
−=
f
fNu
<<
≤≤
63 105Re103
2000Pr5.0
xx ( 3.5)
Cengel [6] stated that, the turbulent flow in circular tube relations could be used for
turbulent flow in noncircular tubes. His explanation on the subject is: “The velocity
and temperature profiles in turbulent flow are nearly straight lines in the core region,
and any significant velocity and temperature gradients occur in the viscous sub
layer.” [6] Figure 3.4 shows the turbulent, overlap and laminar sub layers.
Figure 3.4 Viscous sub layer.
“Despite the small thickness of laminar sub layer (usually much less than 1 percent
of the pipe diameter), the characteristics of the flow in this layer are very important
since they set the stage for flow in the rest of the pipe. Therefore, pressure drop and
heat transfer characteristics of turbulent flow in tubes are dominated by the very thin
viscous sub layer next to the wall surface, and the shape of the core region is not of
much significance. Consequently, the turbulent flow relations given above for
circular tubes can also be used for noncircular tubes with reasonable accuracy by
replacing the diameter D in the evaluation of the Reynolds number by the hydraulic
diameter pAD ch /4= .” [6]
50
3.1.2 Pressure Drop Calculations
In analytical calculations, pressure drop through the cooling channel is also
considered. Depending on the plate fin number, operating point of the fan in use is
changing and as a result, flow rate decreases or increases. If the equation of the fan
performance curve is known, it is possible to calculate the volumetric flow rate
against the known pressure value.
Fan performance curve is known, and given in the Appendix B. Volumetric flow rate
(VFR) values against the pressure values are determined from the fan performance
curve and tabulated (Table 3.1).
Table 3.1 Fan Curve Data (DMAP).
Volumetric Flow Rate [CFM] Pressure [Inches of Water] 0.0 2.72 2.5 2.56 5.0 2.36 7.5 2.20
10.0 2.04 12.5 1.96 15.0 1.94 16.0 1.96 18.0 2.00 20.0 2.06 22.5 2.16 25.0 2.24 27.5 2.27 28.5 2.28 30.0 2.24 32.5 2.10 34.0 2.00 37.5 1.76 41.5 1.40 43.5 1.20 47.0 0.80 53.0 0.00
51
By using fan curve data values, performance curve is plotted again changing the axes
positions as VFR (y axis) and Pressure (x axis) (Figure 3.5). The aim of converting
the axes of the fan performance curve is getting a VFR equation dependent on the
pressure values.
Pressure - VFR
0
10
20
30
40
50
60
0.0 0.5 1.0 1.5 2.0 2.5 3.0
Pressure [Inches of Water]
VF
R [
CF
M]
Figure 3.5 Fan Curve Data Plot.
Because it is not possible to add a trend line to a graph given in Figure 3.5, the graph
is divided into three parts and 4th order trend lines are added to the each part as
shown in Figure 3.6. In next steps, according to results of VFR calculations, trend
line selection will be done.
On this new plot, units of VFR and pressure data are converted to m3/s and Pa
respectively. Equations of the trend lines are added also on the graph. Because there
are two cooling channels on the chassis, the VFR values are divided into two. In this
52
case, if the pressure drop is known, VFR can be calculated by using the trend line
equations.
VFR - Pressure
y = -7E-14x4 + 6E-11x
3 - 2E-08x
2 - 4E-06x + 0.0125
y = 2E-10x4 - 4E-07x
3 + 0.0003x
2 - 0.0912x + 11.568
y = 9E-12x4 - 2E-08x
3 + 2E-05x
2 - 0.008x + 1.1956
0
0.0021
0.0042
0.0063
0.0084
0.0105
0.0126
0.0147
0 100 200 300 400 500 600 700 800
Pressure [Pa]
VF
R [
m3
/s]
Figure 3.6 Trend Line Added Fan Curve Plot (DMAP).
Pressure drop through the plate fins can be calculated by using the Eqn. 3.6. Before
starting to the calculations, an initial velocity value is defined and this value is
updated in a loop with iteration.
2
2
14 VKK
D
LfP ec
h
app ρ
++=∆ ( 3.6)
175.02.0Re0962.0
−
−⋅=h
c
appD
Lf ( 3.7)
53
2
4.08.0
−=
p
aK c ( 3.8)
−
−=
p
a
p
aK c 4.01
2
( 3.9)
The calculated pressure drop is the result of plate fins. In addition, pressure drop of
the chassis without any plate fin must be known. In this instance, a simple flow
analysis is done to find the pressure drop of the chassis.
Two different pressure drop values are obtained. One of them is analytically
calculated pressure drop value for plate fins and the other one belongs to the chassis,
which does not have any plate fins. Both of them are summed up and a total pressure
drop value is obtained for the chassis with plate fins. By using the trend line
equations, VFR value is calculated corresponding the total pressure drop. The
calculated VFR value is used for thermal calculations.
3.1.3 Thermal Model Generation and Solutions
The DMAP has different operating conditions, which are given in MIL-STD-810F as
25 ºC and 55 ºC [7]. The DMAP also must also be operated in negative temperature
values up to -40 ºC. Since there is no need for cooling in minus temperatures, only
25 ºC, and 55 ºC operating conditions are considered and calculations are done for
these two conditions.
Calculations are done by using Mathcad software and the code is given in Appendix
C. Appropriate number of plate fins are selected for the necessary cooling capacity
54
by considering both ambient 25 ºC and 55 ºC operating conditions. Analytical
calculations are done according to given constraint.
Most critical input for the analytical calculations is wall temperatures. Experimental
studies done in Aselsan Inc., results of which are represented in Chapter 2, show that
if the average wall temperature of the chassis is about 63.3 °C at 55 °C operating
condition, the CPUs of the electronic processor cards do not reach to the critical
junction temperature. Plate fin detail must be selected for the critical operating
condition 55 °C at which the major cooling problem occurs. Although cooling is not
critical for 25 °C operating condition, analytical calculations are done for this
condition in order to verify the mathematical model twice. Therefore, 33.3 °C wall
temperature of the 25 °C operating condition is used for 25 °C calculations.
3.1.3.1 25 ºC Operating Condition Calculations
25 ºC ambient temperature is one of the operating conditions for the DMAP. Air
properties are given in Table 3.2 are taken from Incropera and DeWitt [8] and
interpolated to 298 K.
Table 3.2 Air properties at 25 ºC.
Cpair (J/kgK) 1007
ρair (kg/m3) 1.1707
µair (kg/ms) 1.836x10-5
Pr 0.708
kair (W/mK) 0.0261
55
Steps of an example calculation are explained here for the number of plate fins 15.
Because the pressure drop of the chassis without any plate fin is needed, by using
Icepak, a flow analysis is done. The result is used in the analytical model to calculate
the total pressure drop values for different plate fin numbers. Fan operating point
gives the pressure drop value and is obtained as 260 Pa. By using the first trend line
equation of Figure 3.6, total pressure drop and corresponding velocity is calculated
with the initial velocity 10 m/s.
Calculation is done by iteration and finally updated velocity value is 7.7 m/s for the
number of fins=15. Pressure drop along the 15 fins is calculated as 61.6 Pa.
Therefore, calculated VFR value is also updated. The last updated VFR value is
0.01034 m3/s. The VFR value is in the range of first trend line equation. Therefore,
first trend line selection for the VFR calculation is confirmed. After calculations of
VFR, thermal calculations are done. Dimensions of the cooling channel are given in
Table 3.3.
Table 3.3 DMAP channel dimensions.
Channel height (m) h 0.158
Channel length (m) L 0.3
Channel width (m) w 0.01
Fin thickness (m) b 0.0016
“n” shows the number of fins and n = 10..40.
“a” shows the distance between two plate fins and can be calculated as below:
1+
⋅−=
n
bnha ( 3.10)
56
Wetted perimeter of cooling channel,
2)( ⋅+= awP ( 3.11)
Total perimeter of all cooling channels,
PnPall ⋅+= )1( ( 3.12)
Total convection surface area,
LPA allall ⋅= ( 3.13)
Section flow area of one channel
awAs ⋅= ( 3.14)
All section flow area of channels
)1( +⋅= nAA ssall ( 3.15)
Hydraulic diameter cooling channel,
P
AD s
H ⋅= 4 ( 3.16)
Velocity of the cooling air in the channel,
sall
air
airA
VFRV = ( 3.17)
Reynolds number formulation
air
Hairair DV
µ
ρ ⋅⋅=Re ( 3.18)
57
Now friction factor and can Nusselt number be calculated by using Petukhov
correlation:
2)64.1log(Re)82.1(
1
−⋅=f ( 3.19)
( )
−⋅
⋅+
⋅−⋅
=
1Pr8
7.121
Pr1000Re8
3
25.0f
f
Nu ( 3.20)
Most critical parameters for the calculations are Vair and Re. For the selected number
of fins, these two parameters can be calculated. Indeed if the number of fins
increases, pressure drop also increases and operating point of fan must be changed.
Therefore, pressure drop calculations are also done.
After the calculation of Nusselt number, convection heat transfer coefficient can be
calculated by using the Eqn. 3.21:
NuD
kh
H
air ⋅= ( 3.21)
In order to find a result closer to experimental studies, fin efficiencies also must be
calculated. The calculation terms of fin efficiency are given below:
( )( )
2
1
2
⋅⋅
+⋅=
Lbk
Lbhme
al
( 3.22)
( )wme
wmef
⋅
⋅=
tanhη ( 3.23)
58
2⋅⋅= LwA f ( 3.24)
( )( )
f
all
f
oA
Anηη −⋅
⋅+−= 1
11 ( 3.25)
Operating condition temperature is the inlet temperature of cooling air. For the
cooling air outlet temperature, 0.1 ºC difference initial value is assumed and finally
the value is updated with iteration. The critical temperature values used in the
calculations are given in Table 3.4.
Table 3.4 Critical temperatures of 25 ºC operating condition.
Tw (Wall temperature K) 306.5
Ti (Cooling air inlet temperature K) 298
Te (Cooling air outlet temperature K) 301.8
It is necessary to calculate the logarithmic temperature difference for the heat
dissipation calculation.
ewe TTT −=∆ ( 3.26)
iwi TTT −=∆ ( 3.27)
Logarithmic temperature difference (LTD) formulation:
∆
∆
∆−∆=∆
i
e
ie
T
T
TTT
lnln ( 3.28)
59
By using the above given definitions, calculations are done for the LTD and
efficiency. Results are given in Table 3.5.
Table 3.5 LTD and efficiency of channel at 25 ºC.
∆Tln (K) 6.2
me (1/m) 19
ηf 0.988
ηo 0.994
Finally, total heat dissipation rate of the plate fins can be calculated from:
( )lnTAhQ allocal ∆⋅⋅⋅= η ( 3.29)
However, convection heat transfer calculations are done, it is necessary also to see
the cooling air heat transfer capacities of the cooling channels.
In order to calculate the heat transfer capacity of the channel cooling air, Eqn. 3.30 is
used:
)( ieairaircapat TTCpMFRQ −⋅⋅= ( 3.30)
Heat dissipation capacities of cooling air and convection heat transfer must be same.
Therefore, a new temperature difference value (∆Tn) is calculated by using Eqn. 3.31.
airair
cal
nCpMFR
QT
⋅=∆ ( 3.31)
60
The assumed exit temperature value is iterated by using new ∆Tn. Therefore, Qcapat is
compensated to the value Qcal with small errors.
First mathematical model is solved for the 15 number of plate fins in order to show
an example calculation step. Results are tabulated and given in Table 3.6. In the next
step, by using the given definitions, calculations are done to find the appropriate fin
configuration.
Table 3.6 Calculation results at 25 ºC for the number of plate fins 15.
Operating Condition (ºC) 25
Number of fins 15
a (mm) 8.4
∆P (Pa) 61.6
Vair (m/s) 7.7
VFR (cfm) 22
Re 4502
Nu 15.1
h (W/m2K) 43.4
∆T (K) 3.9
ƞo 0.9935
Qcal (W) 47.7
For the next study, all calculations are done in order to find the appropriate number
of fins to dissipate the given heat load. Below terms are calculated for the fin number
range 10-40.
� Distance between two plate fins (a),
� Pressure drop (∆P),
� Velocity values for each number of plate fin (Vair),
� Volumetric flow rate (VFR),
� Reynolds numbers (Re),
� Nusselt number (Nu),
61
� Convection heat transfer coefficient (h),
� Cooling air inlet and exit temperature difference (∆T),
� Fin efficiency (η0),
� Calculated heat dissipation rate (Qcal),
The total generated heat inside the chassis is calculated as approximately 100 W in
the Chapter 1, Page 11. Therefore, 50 W heat load must be dissipated via two
symmetric cooling channels. In order to achieve this goal, iterations are done on the
mathematical model and 17 number of plate fins are selected.
Calculation results for 17 number of plate fins are tabulated and given in Table 3.7.
Table 3.7 Calculation results at 25 ºC for the number of plate fins 17.
Operating Condition (ºC) 25
Number of fins 17
a (mm) 7.3
∆P (Pa) 69.7
Vair (m/s) 7.9
VFR (cfm) 21.9
Re 4234
Nu 14.3
h (W/m2K) 44.3
∆T (K) 4.1
ƞo 0.993
Qcal (W) 50.4
Same calculations are performed for 55 ºC operating condition and appropriate
number of fin is decided.
62
3.1.3.2 55 ºC Operating Condition Calculations
Formulations used for the 25 ºC calculations are applicable for the 55 ºC calculations
by using the properties of air given in Table 3.8, which are valid for 55 ºC. Besides
the air properties, cooling air inlet and wall temperatures are also different. Similarly
0.1 ºC initial temperature difference is assumed between inlet and exhaust air.
Critical wall temperature is also known for the 55 ºC ambient conditions.
Table 3.8 Air properties at 55 ºC.
Cpair (J/kgK) 1008.1
ρair (kg/m3) 1.0682
µair (kg/ms) 1.978x10-5
Pr 0.703
kair (W/mK) 0.0284
The critical temperature of electronic cards components is approximately 95 ºC and
known that if the wall temperature of the chassis is approximate 63.3 ºC, components
of the electronic cards do not reach to the critical temperature. Actually, this is the
design criterion. In the analytical calculations, this critical wall temperature is used.
Therefore, the calculations start with the wall temperature 336.5 K. The critical
temperature values used in the calculations are given in Table 3.9.
Table 3.9 Critical temperatures of 55 ºC operating condition
Tw (Wall temperature K) 336.5
Ti (Cooling air inlet temperature K) 328
Te (Cooling air outlet temperature K) 333
63
The same calculations to find the appropriate number of fin to dissipate 100 W heat
load are done and results are tabulated (Table 3.10) for the 55 ºC operating
conditions. Iterations are done and to dissipate 50 W heat load via each channel,
number of plate fins is found as 21.
Table 3.10 Calculation results at 55 ºC for the number of plate fins 21.
Operating Condition (ºC) 55
Number of fins 21
a (mm) 5.7
∆P (Pa) 82.7
Vair (m/s) 8.2
VFR (cfm) 21.7
Re 3205
Nu 10.8
h (W/m2K) 42.3
∆T (K) 4.6
ƞo 0.9926
Qcal (W) 50.7
3.1.3.3 25 ºC and 55 ºC Fin Number Comparison
There are totally 100 W heat loads to dissipate and there are 2 different operating
conditions. The calculations are done for both 25 ºC and 55 ºC operating conditions
in order to find the necessary plate fin numbers. After the calculations are done the
values are tabulated and given in Table 3.11.
Table 3.11 25 ºC and 55 ºC results comparison.
Operating Conditions ºC
Necessary fin number to dissipate 50 W
25 17
55 21
64
If the both operating conditions compared it is seen that more fin number is
necessary for the 55 ºC condition. Therefore, 55 ºC condition can be taken as the
worst case and fin number selection must be done for this case. Therefore, 21 plate
fin numbers must be used on the cooling channels.
Actually, it may be possible to dissipate more heat loads by using much more plate
fin numbers. However, using more fin numbers than necessary generates useless
mass load on the chassis. Because this chassis is helicopter equipment, mass is an
important criterion. On the other hand, it is not easy to produce plate fins with small
pitches. Chassis having more plate fins increases the cost of the production.
All calculation results of 25 ºC and 55 ºC operating conditions tabulated and given in
Appendix D. The plotted results are in Figure 3.7. Selection of number of fins for a
new chassis with similar dimensions can be done by using the quick reference tables
or for a different size chassis TMT is applicable.
Number of Fin - Qcal
0
10
20
30
40
50
60
70
80
0 5 10 15 20 25 30 35 40 45 50 55
Number of Fin
Qca
l [W
]
25ºC
55ºC
Figure 3.7 Results of TMT for the operating conditions 25 ºC and 55 ºC.
65
While all cooling calculations are done, assumed that all the generated heat is
dissipated by convection. In order words, the amount of dissipated heat by
conduction and radiation can be neglected. In order to support conduction negligence
below explanation is useful:
Chassis surfaces do not have any contact with anywhere. Only bottom surface of the
chassis has a contact with the mounting tray but the mounting tray is mounted to the
ground with elastic isolators. Therefore, there is not any contact with anywhere for
the conduction cooling. In order to see if the negligence of the radiation is true,
radiation heat transfer is also checked.
3.1.4 Radiative Heat Transfer Calculations
In this section, radiative heat transfer calculations are done. Results are examined to
understand if neglecting radiave heat transfer is meaningful for this study. In order to
simulate the real test case, a cabinet is formed which is used instead of test oven and
a block is placed inside the oven, which is substituting the chassis.
Because front surface of chassis is used as a connector plate and the back surface is
used for cooling air output, radiative heat transfer from these two planes are
neglected since these plates do not have a considerable surface area. These two
surfaces are shown in Figure 3.8.
In addition, bottom surface of the chassis is in perfect contact with the oven surface
and thus only 3 walls (2 side surfaces and 1 top surface) of chassis are considered for
radiative heat transfer calculations. To calculate radiative heat transfer from
considered walls, eqn. 3.32 is used.
,11
11bj
N
j
jibioijji
N
j ji
i EFEHqFq
∑∑=
−−
=
−=+
−−
εε .,...,2,1 Ni = ( 3.32)
66
jiF − term is for the view factors between the walls of the whole system. Wall
numbers are shown in Figure 3.9.
Figure 3.8 General view of DMAP in the chamber
Front Surface Back Surface
67
Figure 3.9 Walls numbers.
To describe wall numbers clearly, Table 3.12 is given.
Table 3.12 Wall Numbers.
Wall Descriptions Wall Number
DMAP body minz 1
DMAP body maxy 2
DMAP body maxz 3
OVEN body maxx 4
OVEN body maxy 5
OVEN body maxz 6
OVEN body minx 7
OVEN body miny 8
OVEN body minz 9
5 - Top wall
4 - Rear wall
7 - Front wall
8 – Bottom wall
6 – Right wall 9 – Left wall
1 – DMAP Left Wall
2 – DMAP Top Wall
3 – DMAP right Wall
Oven
DMAP
68
To calculate view factors it is not easy to use a view factor catalogue because of 3D
geometry. View factor catalogues are useful for simple 2D planes. Therefore, Icepak
is used to get view factor values. The geometry is defined and Icepak calculates view
factors. An example result is given in Figure 3.10. These view factor values are
exported and Table 3.13 is formed. These values are used as jiF − in Eqn. 3.32. There
are 9 unknown radiative heat transfer values as 987654321 ,,,,,,,, qqqqqqqqq . Eqn.
3.32 must be written for all 9 walls. At the end of solution all radiative heat transfer
values of all walls are obtained.
Figure 3.10 View factors.
0.2 emissivity value can be used for oven walls ( 987654 ,,,,, εεεεεε ) and 0.3
emissivity value is convenient for chassis walls ( 321 ,, εεε ). Because there is no
external radiation there is no oiH term in the equation. Equations are written in
matrix form and Mathcad is used to solve the equations.
69
Table 3.13 View factors.
F12 0.00000 F21 0.00000 F31 0.00000 F41 0.00371 F51 0.00145
F13 0.00000 F23 0.00000 F32 0.00000 F42 0.00666 F52 0.01171
F14 0.07074 F24 0.17293 F34 0.07066 F43 0.00371 F53 0.00145
F15 0.02764 F25 0.30409 F35 0.02761 F45 0.19988 F54 0.19988
F16 0.00000 F26 0.17131 F36 0.42392 F46 0.19954 F56 0.19991
F17 0.07478 F27 0.18037 F37 0.07478 F47 0.19505 F57 0.19976
F18 0.40303 F28 0.00000 F38 0.40303 F48 0.18272 F58 0.18539
F19 0.42382 F29 0.17131 F39 0.00000 F49 0.19958 F59 0.19981
F61 0.00000 F71 0.00392 F81 0.02199 F91 0.02223
F62 0.00660 F72 0.00695 F82 0.00000 F92 0.00660
F63 0.02224 F73 0.00392 F83 0.02199 F93 0.00000
F64 0.19954 F74 0.19505 F84 0.19004 F94 0.19958
F65 0.19991 F75 0.19976 F85 0.19282 F95 0.19981
F67 0.19952 F76 0.19952 F86 0.18461 F96 0.19295
F68 0.17750 F78 0.18199 F87 0.18928 F97 0.19948
F69 0.19295 F79 0.19948 F89 0.18473 F98 0.17762
There are 3 matrixes. A is the unknown “q” matrix. B is the coefficient matrix of
unknown “q” values. Finally C is the known values matrix. All matrixes are shown
below. According to Eqn. 2.27, CBA =* . Because it is wanted to find unknown “q”
values, the matrix equation CBA *1−= is used. The critical input for the
calculations is wall temperature values and is the result of experiments. The
calculations are done for the operating temperature 25°C. So all chassis walls are at
306.3K and oven walls are at 298K.
By using Mathcad the matrix multiplication is done.
70
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
⋅−⋅−⋅−⋅−⋅−⋅−⋅−⋅−
=
8987976965954943932921919
9897876865854843832821818
9798786765754743732721717
9698687675654643632621616
9598587576564543532521515
9498487476465454432421414
9398387376365354342321313
9298287276265254243231212
9198187176165154143132121
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
EbFEbFEbFEbFEbFEbFEbFEbFEb
C
=
9
8
7
6
5
4
3
2
1
q
q
q
q
q
q
q
q
q
A
2.0
2.0
2.0
2.0
2.0
2.0
3.0
3.0
3.0
9
8
7
6
5
4
3
2
1
=
=
=
=
=
=
=
=
=
ε
ε
ε
ε
ε
ε
ε
ε
ε
KT
KT
KT
KT
KT
KT
KT
KT
KT
298
298
298
298
298
298
3.306
3.306
3.306
9
8
7
6
5
4
3
2
1
=
=
=
=
=
=
=
=
=
4
99
4
88
4
77
4
66
4
55
4
44
4
33
4
22
4
11
TEb
TEb
TEb
TEb
TEb
TEb
TEb
TEb
TEb
⋅=
⋅=
⋅=
⋅=
⋅=
⋅=
⋅=
⋅=
⋅=
σ
σ
σ
σ
σ
σ
σ
σ
σ
42
81067.5Km
W⋅×= −σ
71
72
Mathcad result is shown below. To calculate all radiation heat transfer values we
must multiply below “q” results by area of walls.
2
618.1
577.2
664.1
618.1
564.1
21.2
69.11
285.11
688.11
m
WA ⋅
= 2
3
2
1
/
7.11
3.11
7.11
mW
q
q
q
≈
≈
≈
Area of wall 1 is = 052.0332.0158.0 =x m2
Area of wall 2 is = 039.0332.0116.0 =x m2
Area of wall 3 is = 052.0332.0158.0 =x m2
The radiative heat transfer results of chassis walls are given below. The value Q1, Q2
and Q3 are the radiative heat dissipation rates of surfaces 1, 2 and 3 respectively
shown in Figure 3.9.
9.01 ≈Q W
7.02 ≈Q W
9.03 ≈Q W
In conclusion, wall temperatures are the results of experiments. If the radiative heat
transfer calculation results are examined it is seen that totally 2.5 W heat rate is
dissipated with radiation. If the radiation and convection results are examined
together, it seen that 2.5 W can be neglected compared to 100 W total load.
Therefore, radiative heat transfer neglect is meaningful while analytical calculations
are done.
73
3.2 ACCC Analytical Studies
ACCC is the existing chassis on which experimental studies are performed. The
purpose of the analytical studies on ACCC is to indicate the applicability of TMT on
different chassis. The TMT is developed on DMAP but can be applied on any type of
ATR chassis by changing specific parameters as dimensions and trend line
equation of the used fan. Number of plate fins=24 was used in the design of ACCC
and the chassis was produced in this wise. Therefore, dissipated heat is calculated by
using TMT for the number of fins 24 for both the operating conditions 25 ºC and 55
ºC. Similar to the DMAP studies, in the next step numerical studies are done for
ACCC and results are compared.
Ametek Rotron Propimax 2 fan shown in Figure 3.11 is used in ACCC. Fan
performance curve and dimensions of the fan is given in Appendix B. VFR –
Pressure plot is needed to get a fan curve dimension. Therefore, fan curve data is
acquired on the fan performance curve. VFR (y axis) and Pressure (x axis) graph is
plotted by converting VFR and pressure units to m3/s and Pa respectively (Figure
3.12). A 4th order trend line with equation is added on the curve.
Figure 3.11 Propimax 2 Fan.
74
Table 3.14 Fan Curve Data (DMAP).
Volumetric Flow Rate [CFM]
Pressure [Inches of Water]
0 3.2
7.1 3
20 2.6
40 1.9
60 1.4
80 1.1
100 0.7
118.6 0
VFR - Pressure
y = -5E-13x4 + 9E-10x
3 - 5E-07x
2 + 1E-05x + 0.056
0.000
0.010
0.020
0.030
0.040
0.050
0.060
0 100 200 300 400 500 600 700 800 900
Pressure [Pa]
VF
R [
m3/s
]
Figure 3.12 Trend Line Added Fan Curve Plot (ACCC).
75
Pressure drop value of the ACCC without any plate fins is required to use in TMT.
Therefore, a quick CFD flow analysis is performed and 435 Pa pressure drop is
obtained. Calculations are made for both 25 ºC and 55 ºC operating conditions using
trend line equation, pressure drop and dimensions of ACCC (Table 3.15) and results
are tabulated and given in Table 3.16. Numerical studies will be performed using the
same inputs and results will be compared.
Table 3.15 ACCC Channel Dimensions.
Channel height (m) h 0.192
Channel length (m) L 0.30
Channel width (m) w 0.0068
Fin thickness (m) b 0.0016
Table 3.16 Calculation results for 24 number of plate fins (ACCC).
Operating Condition (ºC) 55 25
Number of fins 24
a (mm) 6.1
∆P (Pa) 165.2 160.4
Vair (m/s) 11.1 11.3
VFR (cfm) 24.5 25
Re 4564 3945
Nu 15.3 13.3
h (W/m2K) 62 58.7
∆T (K) 5 5
ƞo 0.9959 0.9961
Qcal (W) 67.5 63.5
76
CHAPTER 4
4 NUMERICAL STUDIES
Numerical studies are conducted to verify analytical calculations. In the analytical
calculations chapter, necessary number of fin was selected to dissipate the generated
heat loads for the chassis DMAP. This chapter involves numerical investigations of
the designed chassis. Numerical studies performed using Icepak 4.4.8 that is a CFD
software that solves mass, momentum and energy conservation equations to simulate
fluid flows with heat transfer. Radiative transport equation is not included in the
computations because of the negligence of radiative heat transfer in analytical
calculations.
Numerical studies are carried out for the chassis DMAP and ACCC both for 25 ºC
and 55 ºC operating conditions. Steps of the numerical studies can be listed as below:
• 3D model creation: The geometry is modeled in Icepak with the same
constraints used in analytical studies.
• Meshing: The created model is meshed using the right meshing parameters
with the appropriate meshing procedure.
• Solution: The meshed model is solved using the right solution procedure and
results are examined.
77
4.1 Numerical Studies of DMAP
At the first step, the geometry is modeled in the Icepak with the known constraints,
which are used in analytical calculations. Cooling channels are created on the side of
the chassis and plate fins are placed in the cooling channels. There is a cabinet
geometry, which surrounds the chassis. The general view of the created 3D model is
shown in Figure 4.1.
The fan is placed at the back cover of the chassis as in the real model. Ametek
Rotron Vaneaxial Aximax 2 model fan is used in the design. The fan performance
curve (shown in Appendix A) is used during the analysis.
In order to reduce the mesh size and analysis time inside the chassis is defined as
hollow block shown in Figure 4.1. Therefore, as in the mathematical model only
cooling channels of the chassis are analyzed.
Wall temperature values were used as inputs for analytical calculations and necessary
number of plate fns are calculated for the given heat load. In numerical calculations
dissipated heat, which is the result of TMT, defined as total power on the sidewalls
and wall temperatures are found numerically.
78
Figure 4.1 3D Model of DMAP.
4.1.1 Boundary Conditions and Basic Parameters of DMAP
Icepak automatically creates a cabinet to use as computational domain shown in
Figure 4.2 with blue boundary lines. The cabinet dimensions are modified to 116 x
158 x 332 mm in X-Y-Z directions respectively. Because radiation is not included,
there is no clearance between cabinet and DMAP walls. Except sidewalls, other four
walls of the cabinet are adiabatic. Because in the analytical calculations, studies are
conducted only for the cooling channels. Sidewalls are defined as opening and open
to mass and heat transfer. However, sidewalls of the chassis are defined adiabatic
Plate Fins
Cooling Channels
Fan
Hollow Block
79
except front air intake openings shown in Figure 4.2. The adiabatic sidewall of the
chassis is hidden in the Figure 4.2 to show the plate fins. The analytically calculated
50 W heat load is defined on the sidewalls as boundary conditions. Inlet sections are
defined as openings for the boundary conditions with the ambient temperature. Fan is
the outlet boundary for the chassis.
Figure 4.2 Computational Model of DMAP.
In the numerical calculations, flow and temperature equations are solved and
radiative heat transfer is neglected. Because the velocity and Re is calculated in
analytical studies about 7.9 m/s and 4200 respectively, turbulent two-equation flow
regime is selected. Ambient temperature is defined as 25 ºC for the first and 55 ºC for
the second simulation. Because the problem is not time dependent, time variation is
chosen as steady.
Cabinet
Opening (On both sides)
80
4.1.2 Grid Generation on DMAP
There are three types of meshers available in ANSYS Icepak: hexahedral, tetrahedral
and hex-dominant. The hexahedral unstructured mesher (the default for most
applications) is widely used and the most appropriate mesher [22]. Because the
geometry is not complicated and does not include spherical or ellipsoidal objects, the
hexahedral mesher is selected. Because mesh generation should be an iterative
procedure for best results, different numerical studies are done to find the true mesh
size changing the max size ratios. Therefore, mesh independence of the solution is
also verified.
For the fist case (number of plate fins 17), DMAP is meshed using mesh control
parameters given in Figure 4.4. By using per-object mesh parameters given, element
height of the fin is controlled and defined as 0.1 mm for refinement. 1,247,124
number of elements are generated shown in Figure 4.3 and quality ranges of the
elements are shown in Figure 4.6. Detailed view of a meshed fin is shown in Figure
4.5. Different numerical studies are performed for the max size ratios 2, 1.7 and 1.5.
81
Figure 4.3 Meshed view of the chassis.
Figure 4.4 Mesh Control Parameters.
82
Figure 4.5 Detailed view of meshed fin.
Figure 4.6 Quality of the mesh.
83
4.1.3 Numerical Solution of DMAP
First-order discretization scheme is selected for the temperature and momentum.
Then by default, pressure equation is solved using the standard scheme, which gives
a relatively quick and accurate solution [19,22]. Under-relaxation factors and linear
solver settings are used as default. The version of the solver is selected as single
precision since the geometry does not have very disparate length scales [22].
Advanced solver setup window is shown in Figure 4.7. Turbulent, two-equation
model is used for the solution.
Figure 4.7 Advanced solver setup window.
84
4.1.3.1 25 ºC Operating Condition Solution of DMAP
25 ºC analysis is done to verify the TMT results of 25 ºC. Therefore; the chassis has
number of plate fins 17 for this analysis. Operating condition is selected as 25 ºC and
50W heat load is defined on each wall as the design criteria. Analysis is completed
approximately after 1500 iterations. In order to understand if the convergence occurs;
points are placed on different zones of the chassis. Velocity, pressure, and
temperature curves of the points are followed if the change in the values stopped.
The points with the name W-1, W-2 and W-3 are on the wall and the point with the
name C is in the channel.
This study is repeated for three different mesh options, which are generated changing
max size units as 2, 1.7 and 1.5. The results of max size ratio 2 are given in detail and
the results of other size ratios are given in Table 4.1. ∆T shows the temperature
difference of cooling air between inlet and outlet sections. Qcal terms are calculated
using the VFR and �T variables and air properties with the heat transfer capacity
equation TCmQ ∆⋅⋅= .
Table 4.1 25 ºC results of DMAP for different number of elements.
Max Size ratio 1.5 1.7 2
Number of elements 2803737 1652596 1247124
∆P (Pa) 78.8 76.2 76.5
Vair (m/s) 7.2 7.2 7.1
VFR (CFM) 20.1 19.9 19.5
∆T (ºC) 4.9 4.9 4.9
Wall Temp. (ºC) 30.2 30.3 30.3
Qcal (W) 54 53 51
Because the results are nearly the same, it is possible to say that the analyses are
independent of mesh sizes. If the max size ratio is too high and as a result if the
overall mesh is too coarse, the resulting solution may be inaccurate. If the max size
ratio is too low and as a result if the overall mesh is too fine, the computational cost
85
may become prohibitive. In summary, the cost and accuracy of the solution are
directly dependent on the quality and size of the mesh [22]. All terms given in Table
4.1 are the direct results of the simulations except Qcal term. The simulation result
with the max size ratio 2 is believed to be most accurate. Because the Qcal output of
this simulation is closest to the defined 50 W heat load on the wall. Therefore, next
grids are generated for all simulations using the max size ratio 2, which has also
minimum element size.
During analytical studies, calculations are done using the same temperature values
along the fins. To verify this, also 6 different points are placed on one of the plate
fin, which is in the left cooling channel. Detailed view of these points is shown in
Figure 4.8. As shown temperature values are very close to each other along the fin.
Figure 4.8 Control points.
86
Wall temperature results are given in Table 4.2. The average temperature of the wall
is calculated as 30.2 °C using the points on the wall. In addition, mean temperature
value of the wall, shown in Figure 4.9, is 30.3 °C. In the analytical calculations 33.3
°C values is used for the wall temperature.
Table 4.2 Temperature results of points at 25 ºC operating condition (DMAP).
W-1 W-2 W-3
29.1 ºC 30.3 ºC 31.3 ºC
Average: 30.2 ºC
Figure 4.9 Wall temperature contour view of 25 ºC analyses (DMAP).
Numerical study results of 25 °C are given in Table 4.3 with the results of 25 °C
analytical studies.
Table 4.3 Results of points at 25 ºC operating condition (DMAP).
25 ºC (17 Fin)
Analytical Numerical
∆P (Pa) 69.7 76.5
Vair (m/s) 7.9 7.1
VFR (cfm) 21.9 19.5
∆T (K) 4.1 4.9
Wall Temp. (ºC) 33.3 30.3
Qcal (W) 50.4 51
87
4.1.3.2 55º Operating Condition Solution of DMAP
Same analysis is done for the 55 °C operating condition and results are tabulated in
Table 4.4. In the analytical calculations 63.3 ºC average wall temperature is used for
55 ºC operating conditions. If analysis results are examined, average wall
temperature is shown as 59.9 ºC. In addition, mean temperature value of the wall,
shown in Figure 4.10, is 60 °C.
Table 4.4 Temperature results of points at 55 ºC operating condition (DMAP).
W-1 W-2 W-3
58.7 ºC 59.9 ºC 61.1 ºC
Average: 59.9 ºC
Figure 4.10 Wall temperature contour view of 55 ºC analyses (DMAP).
Numerical study results of 55 °C are given in Table 4.4 with the results of 55 °C
analytical studies.
88
Table 4.5 Temperature results of points at 25 ºC operating condition (DMAP).
55 ºC (21 Fin)
Analytical Numerical
∆P (Pa) 82.3 110.4
Vair (m/s) 8.2 7.8
VFR (cfm) 21.7 21
∆T (K) 4.6 5.1
Wall Temp. (ºC) 63.3 60
Qcal (W) 50.7 51
4.2 Numerical Studies of ACCC
Same procedures of numerical studies of DMAP given in section 4.1 are followed for
the numerical studies of ACCC. The general view of the created 3D model is shown
in Figure 4.1. Similar to DMAP, to reduce the mesh size and analysis time inside the
chassis is defined as hollow block.
The fan is placed at the back cover of the chassis as in the real model. Ametek
Rotron Propimax 2 model fan is used in the design. The curve of the fan, which is
given in Appendix B, is used during the analysis.
89
Figure 4.11 3D Model of ACCC.
Figure 4.12 3D Model of ACCC (continued).
Hollow Block
Air intake
Air exhaust
Cooling Channels
Plate Fins
Fans
Cabinet Boundary
90
4.2.1 Boundary Conditions and Basic Parameters of ACCC
The generated cabinet dimensions are 257.8 x 192 x 380 mm in X-Y-Z directions
respectively. Air intake and exhaust sections of ACCC are at the front and back sides
as shown in Figure 4.11 and Figure 4.12. Therefore, front and back walls of the
cabinet are defined as openings; other four walls of the cabinet are adiabatic. Basic
parameters of the solution are same with the DMAP. Turbulent two-equation flow
regime is selected. Ambient temperature is defined as 25 ºC for the first simulation
and 55 ºC for the second simulation. Time variation is chosen as steady.
4.2.2 Grid Generation on ACCC
ACCC is meshed using the same mesh control parameters given in section 4.1.2.
1,727,574 number of elements are generated shown in Figure 4.13.
Figure 4.13 Meshed view of the chassis.
91
4.2.3 Numerical Solution of ACCC
Same solution setting is used given in section 4.1.3. First-order discretization scheme
is selected for the temperature and momentum and by default pressure equation is
solved using the standard scheme. Under-relaxation factors and linear solver settings
are used as default. The version of the solver is selected as single precision.
Turbulent, two-equation model is used.
4.2.3.1 25 ºC Operating Condition Solution of ACCC
25 ºC analysis is preformed to verify the TMT results of 25 ºC of ACCC. Operating
condition is selected as 25 ºC. 68 W heat loads is defined on each wall as the result
of TMT solution. Analysis is completed approximately after 1000 iterations. In order
to understand if the convergence is came true; points are placed on different zones of
the chassis. The points with the name W-1, W-2 and W-3 are on the wall and the
point with the name C is in the channel.
92
Figure 4.14 Wall temperature contour view of 25 ºC analyses (ACCC).
The result of the analysis is given inTable 4.6. The average temperature of the wall is
calculated as 30.0°C using the points on the wall. In addition, mean temperature
value of the wall, shown in Figure 4.14, is 30.2°C. In the analytical calculations
33.3°C values is used for the wall temperature.
Table 4.6 Temperature results of points at 25 ºC operating condition (ACCC).
W-1 W-2 W-3
28.9 ºC 30.0 ºC 31.2 ºC
Average: 30.0 ºC
Numerical study results of 25 °C are given in Table 4.7 with the results of 25 °C
analytical studies.
93
Table 4.7 Temperature results of points at 25 ºC operating condition (ACCC).
25 ºC (24 Fin)
Analytical Numerical
∆P (Pa) 165.2 167
Vair (m/s) 11.1 10.8
VFR (cfm) 24.5 23.9
∆T (K) 4.9 5
Wall Temp. (ºC) 33.3 30.4
Qcal (W) 67.5 68
4.2.3.2 55 ºC Operating Condition Solution of ACCC
Same analysis is done for the 55 °C operating condition and results are tabulated in
Table 4.8. In the analytical calculations average 63.5 ºC average wall temperature is
used for 55 ºC operating conditions. If analysis results are examined, average wall
temperature is shown as 59.9 ºC. In addition, mean temperature value of the wall,
shown in Figure 4.10, is 60 °C.
Table 4.8 Temperature results of points at 55 ºC operating condition (ACCC).
W-1 W-2 W-3
58.8 ºC 59.9 ºC 61.0 ºC
59.9 ºC
94
Figure 4.15 Wall temperature contour view of 55 ºC analyses (ACCC).
Numerical study results of 55 °C are given in Table 4.9 with the results of 55 °C
analytical studies.
Table 4.9 Temperature results of points at 25 ºC operating condition.
55 ºC (24 Fin)
Analytical Numerical
∆P (Pa) 160.4 167
Vair (m/s) 11.3 11.8
VFR (cfm) 25 26.2
∆T (K) 5 4.7
Wall Temp. (ºC) 63.3 60
Qcal (W) 63.5 64
95
CHAPTER 5
5 DISCUSSION
The aim of this thesis is to develop a thermal model tool (TMT) for standard Avionic
Transport Rack (ATR) chassis and make the thermal design of a standard ATR
chassis Digital Moving Map (DMAP) using developed TMT.
The results of experimental studies are utilized in analytical studies by considering
mechanical and thermal limitations, subsequently TMT is developed. To double
check the accuracy of TMT, also plate fin details of the existing chassis ACCC
(Avionic Central Control Computer) is determined in analytical studies chapter. In
the next step, numerical verification of TMT is accomplished. Eventually, the
purpose of the thesis study is achieved its goal.
In this section, experimental, analytical, and numerical studies results are compared.
All these studies are done for both 25 ºC and 55 ºC operating conditions however, the
critical operating condition is 55 ºC because of cooling problems at high
temperatures. All studies are based on the critical wall temperatures determined in
experimental studies. At the 55 °C operating condition, average critical wall
temperature is measured as 63.3 °C. This temperature value is the results of
experimental studies of ACCC. When the wall temperature is approximately 63.3 °C
at 55 °C operating condition, junction temperatures of the CPUs are in the safe
operation range. However, if the wall temperature exceeds 63.3 °C, the junction
temperatures of the CPUs are over the safe operating temperature range. Therefore,
in all studies for the operating condition 55 °C, 63.3 °C critical wall temperature is
96
considered. Besides this, average wall temperature is known for 25 °C operating
condition and this is used in order to double check the TMT.
The summary of the experimental studies are given in Table 5.1. Average wall
temperatures are calculated 33.3°C and 63.3°C for the 25 °C and 55 °C operating
conditions respectively.
Table 5.1 Results of experimental studies.
Operating Condition
Slot1 Slot2 Slot3 Slot4 Slot5 Slot6 Slot7 Slot10 Slot11 Average
Steady State Temperatures [°C]
25 33.7 36.6 34.7 34.6 33.8 32.7 31.8 30.8 30.7 33.3
55 63.6 66.7 64.7 64.7 63.7 62.6 61.8 61.0 60.5 63.3
The summary of the analytical and numerical studies is given in Table 5.2 and Table
5.3 for DMAP and ACCC respectively. Calculations are done for both 25 °C and 55
°C operating conditions using TMT.
Table 5.2 Analytical and numerical results of DMAP.
DMAP
25 ºC (17 Fin) 55 ºC (21 Fin)
Analytical Numerical Analytical Numerical
∆P (Pa) 69.7 78.8 82.3 110.4
Vair (m/s) 7.9 7.1 8.2 7.8
VFR (cfm) 21.9 20 21.7 21
∆T (K) 4.1 4.9 4.6 5.1
Wall Temp. (ºC) 33.3 30.3 63.3 60
Qcal (W) 50.4 51 50.7 51
97
Table 5.3 Analytical and numerical studies results of ACCC.
ACCC
25 ºC (24 Fin) 55 ºC (24 Fin)
Analytical Numerical Analytical Numerical
∆P (Pa) 165.2 167 160.4 167
Vair (m/s) 11.1 10.8 11.3 11.8
VFR (cfm) 24.5 23.9 25 26.2
∆T (K) 4.9 5 5 4.7
Wall Temp. (ºC) 33.3 30.4 63.3 60
Qcal (W) 67.5 68 63.5 64
Four different case studies are conducted both analytically and numerically. It is
observed that the results are close. The generated TMT is verified numerically both
for DMAP and ACCC. In the next designs, the results of this thesis will be used and
without doing long-term numerical analysis, effective plate fin number will be
determined. By using generated tables for different operating conditions, heat
dissipation capacities of similar size chassis will be determined easily. In addition, by
using the TMT, heat dissipation capacities of different size chassis will be
determined without numerical studies in great detail.
Plate fin numbers are selected for the worst-case 55 °C operating conditions for the
new design chassis DMAP. As shown in Table 5.2, 21 plate fins are sufficient for the
cooling channels to dissipate 50 W heat load via each channel.
In this study;
• Uniform heat distribution on the chassis wall is assumed and only one
thermocouple is placed on each slot.
• Fin temperature is assumed same along the fins.
98
• Symmetry condition is assumed between left and right walls of the chassis.
In the future work;
• Heat distribution maps on the walls can be generated by using a thermal
camera.
• Differences between cooling channels can be determined and numbers of
each cooling channel can be calculated separately.
99
REFERENCES
[1] Aeronautical Radio, Inc., 1974, “Air Transport Equipment Cases and
Racking,” ARINC 404A.
[2] Macrolink Inc., http://www.macrolink.com/CHASSIS/ATR/ml750t.htm, last
visited on April 2011.
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102
APPENDIX A
APENDICES
A. COOLING CHANNEL DIMENSIONS OF DMAP
Figure A.1 Front section view of cooling channels.
103
Figure A.2 Side view of cooling channel.
104
APPENDIX B
B. FAN PERFORMANCE CURVE AND DIMENSIONS
Figure B.1 Fan performance curve of Aximax 2.
105
Figure B.2 Dimensions of Aximax 2.
106
Figure B.3 Fan performance curve of Propimax 2.
107
Figure B.4 Dimensions of Propimax 2.
108
APPENDIX C
C.MATHCAD CODE FOR FIN OPTIMIZATION
n 1 50..:=
CP.air 1007J
kg K⋅:=
ρair 1.1707kg
m3
:=
µair 1.836105−
⋅kg
m s⋅:=
kal 151W
m K⋅:=
kair 0.0261W
m K⋅:=
Pr 0.708:=
L 0.3 m⋅:= Channel lenth
w 0.01m:= Channel width
H 0.158m:= Channel height
b 0.0016m⋅:= Fin thickness
an
H n b⋅−
n 1+:= Distance between two fins
Pn
w an
+( ) 2⋅:= Channel perimeter
Palln
n 1+( ) Pn
⋅:= All channel perimeter
Palln
n 1+( ) Pn
⋅:= All channel perimeter
Aalln
Palln
L⋅:= Total convection surface area
109
DHn
4As
n
Pn
⋅:=
Asn
w an
⋅:= Channel flow section area
Asalln
Asn
n 1+( )⋅:= All channel flow section area
pn
an
b+:=
Kcn0.8 0.4
an
pn
2
−:=
Ken1
an
pn
2
−
0.4an
pn
−:=
110
Hizn
Vairi 0.1m
s←
Rein
ρair Vairi⋅ DHn
⋅
µair←
fn
0.0962 Rein( )
0.2−⋅
L
DHn
0.175−
⋅
←
∆Pn
4 fn
⋅L
DHn
⋅ Kcn+ Ken
+
1
2⋅ ρair⋅ Vairi( )
2⋅
1
Pa←
∆Ptn
260 ∆Pn
+←
VFRairn
7− 1014−
⋅ ∆Ptn( )
4⋅ 6 10
11−⋅ ∆Pt
n( )3
⋅+ 2 108−
⋅ ∆Ptn( )
2⋅− 4 10
6−⋅ ∆Pt
n⋅− 0.0125+
m3
s←
Vairn
VFRairn
Asalln
←
Vairi Vairi 0.001m
s+←
Rein
ρair Vairi⋅ DHn
⋅
µair←
fn
0.0962 Rein( )
0.2−⋅
L
DHn
0.175−
⋅
←
∆Pn
4 fn
⋅L
DHn
⋅ Kcn+ Ken
+
1
2⋅ ρair⋅ Vairi( )
2⋅
1
Pa←
∆Ptn
260 ∆Pn
+←
VFRairn
7− 1014−
⋅ ∆Ptn( )
4⋅ 6 10
11−⋅ ∆Pt
n( )3
⋅+ 2 108−
⋅ ∆Ptn( )
2⋅− 4 10
6−⋅ ∆Pt
n⋅− 0.0125+
m3
s←
Vairn
VFRairn
Asalln
←
Vairi Vairn
− 0.01m
s>while
Vairi
:=
111
Vair Hiz:=
Ren
ρair Vairn
⋅ DHn
⋅
µair:=
fn
0.0962 Ren( )
0.2−⋅
L
DHn
0.175−
⋅
:=
∆Pn
4 fn
⋅L
DHn
⋅ Kcn+ Ken
+
1
2⋅ ρair⋅ Vair
n( )2
⋅
1
Pa:=
VFRairn
Vairn
Asalln
⋅:=
MFRairn
VFRairn
ρair⋅:=
ffn
1
1.82 log Ren( )⋅ 1.64−( )
2:=
nun
ffn
8
Ren
1000−( )⋅ Pr⋅
1 12.7ff
n
8
0.5
⋅ Pr
2
31−
⋅+
:=
hn
kair
DHn
nun
⋅:=
men
hn
2⋅ b L+( )
kal b L⋅( )⋅
1
2
:=
ηfn
tanh men
w⋅( )me
n( ) w⋅:=
Af w L⋅ 2⋅:=
112
ηon
1n 1+( ) Af⋅
Aalln
1 ηfn
−( )⋅−:=
Tw 306.5K:=
Ti 298K:=
DeltaTn
tt 0.1K←
Ten298K tt+←
∆T enTw Ten
−←
∆T i Tw Ti−←
∆T lnn
∆T en∆T i−
ln
∆T en
∆T i
←
Qcaln
ηon
hn
Aalln
⋅ ∆T lnn⋅
⋅←
∆Tnn
Qcaln
MFRairn
CP.air⋅←
tt tt 0.01K+←
Ten298K tt+←
∆T enTw Ten
−←
∆T i Tw Ti−←
∆T lnn
∆T en∆T i−
ln
∆T en
∆T i
←
Qcaln
ηon
hn
Aalln
⋅ ∆T lnn⋅
⋅←
∆Tnn
Qcaln
MFRairn
CP.air⋅←
∆Tnn
tt− 0.05K>while
tt
:=
113
Ten298K DeltaT
n+:=
Qcapatn
MFRairn
CP.air⋅ TenTi−
⋅:=
0 2.5 5 7.5 10 12.5 15 17.5 20 22.5 25 27.5 30 32.5 35 37.5 40 42.5 45 47.5 5020
25
30
35
40
45
50
55
60
65
70
Qcapatn
n
114
APPENDIX D
D. ANALYTICAL RESULTS
25 ºC ANALYTICAL RESULTS
a ∆P V VFR Re Fin
Number [mm] [Pa] [m/s] [CFM] [-]
10 12.9 44.0 7.4 22.3 5.3E+03
11 11.7 47.3 7.5 22.2 5.1E+03
12 10.7 50.7 7.5 22.2 5.0E+03
13 9.8 54.2 7.6 22.1 4.8E+03
14 9.0 57.8 7.7 22.1 4.6E+03
15 8.4 61.6 7.7 22.0 4.5E+03
16 7.8 65.6 7.8 21.9 4.4E+03
17 7.3 69.7 7.9 21.9 4.2E+03
18 6.8 74.0 8.0 21.8 4.1E+03
19 6.4 78.4 8.0 21.7 4.0E+03
20 6.0 83.0 8.1 21.6 3.9E+03
21 5.7 87.8 8.2 21.6 3.8E+03
22 5.3 92.8 8.3 21.5 3.7E+03
23 5.1 98.0 8.3 21.4 3.6E+03
24 4.8 103.4 8.4 21.3 3.5E+03
25 4.5 108.9 8.5 21.2 3.4E+03
26 4.3 114.7 8.6 21.1 3.3E+03
27 4.1 120.7 8.6 21.0 3.2E+03
28 3.9 127.0 8.7 20.9 3.1E+03
29 3.7 133.4 8.8 20.8 3.0E+03
30 3.5 140.1 8.9 20.6 3.0E+03
31 3.4 147.0 8.9 20.5 2.9E+03
32 3.2 154.1 9.0 20.4 2.8E+03
33 3.1 161.4 9.1 20.2 2.7E+03
34 3.0 168.9 9.1 20.0 2.7E+03
35 2.8 176.6 9.2 19.9 2.6E+03
36 2.7 184.6 9.3 19.7 2.5E+03
37 2.6 192.7 9.3 19.5 2.5E+03
38 2.5 200.9 9.4 19.3 2.4E+03
39 2.4 209.4 9.4 19.1 2.3E+03
40 2.3 217.9 9.5 18.9 2.3E+03
115
Nu h ∆T ƞo Qcal Fin
Number [-] [W/m2K] [K] [-] [W]
10 17.6 40.8 3.3 0.9951 40.8
11 17.1 41.3 3.4 0.9948 42.2
12 16.6 41.9 3.5 0.9945 43.7
13 16.1 42.4 3.7 0.9941 45.0
14 15.6 42.9 3.8 0.9938 46.4
15 15.1 43.4 3.9 0.9935 47.7
16 14.7 43.8 4.0 0.9933 49.0
17 14.3 44.3 4.1 0.9930 50.4
18 13.9 44.8 4.3 0.9927 51.7
19 13.5 45.2 4.4 0.9925 52.8
20 13.1 45.7 4.5 0.9922 54.1
21 12.8 46.1 4.6 0.9920 55.2
22 12.4 46.5 4.7 0.9917 56.4
23 12.1 46.9 4.8 0.9915 57.5
24 11.7 47.3 4.9 0.9913 58.5
25 11.4 47.6 5.1 0.9911 59.6
26 11.1 48.0 5.2 0.9908 60.5
27 10.8 48.3 5.3 0.9906 61.4
28 10.5 48.6 5.4 0.9905 62.3
29 10.2 48.9 5.5 0.9903 63.1
30 9.9 49.1 5.6 0.9901 63.8
31 9.6 49.4 5.7 0.9899 64.4
32 9.3 49.5 5.8 0.9898 65.1
33 9.0 49.7 5.8 0.9896 65.6
34 8.7 49.8 5.9 0.9895 66.0
35 8.4 49.8 6.0 0.9894 66.5
36 8.2 49.8 6.1 0.9893 66.7
37 7.9 49.8 6.2 0.9892 67.0
38 7.6 49.7 6.3 0.9892 67.1
39 7.3 49.5 6.3 0.9891 67.2
40 7.0 49.2 6.4 0.9891 67.1
116
55 ºC ANALYTICAL RESULTS
a ∆P V VFR Re Fin
Number [mm] [Pa] [m/s] [CFM] [-]
10 12.9 41.3 7.4 22.3 4.5E+03
11 11.7 44.4 7.5 22.3 4.4E+03
12 10.7 47.6 7.6 22.2 4.2E+03
13 9.8 50.9 7.6 22.2 4.1E+03
14 9.0 54.3 7.7 22.1 3.9E+03
15 8.4 57.9 7.8 22.1 3.8E+03
16 7.8 61.6 7.8 22.0 3.7E+03
17 7.3 65.5 7.9 21.9 3.6E+03
18 6.8 69.5 8.0 21.9 3.5E+03
19 6.4 73.7 8.1 21.8 3.4E+03
20 6.0 78.1 8.1 21.7 3.3E+03
21 5.7 82.7 8.2 21.7 3.2E+03
22 5.3 87.4 8.3 21.6 3.1E+03
23 5.1 92.3 8.4 21.5 3.0E+03
24 4.8 97.5 8.4 21.4 3.0E+03
25 4.5 102.8 8.5 21.3 2.9E+03
26 4.3 108.3 8.6 21.2 2.8E+03
27 4.1 114.0 8.7 21.1 2.7E+03
28 3.9 120.0 8.8 21.0 2.7E+03
29 3.7 126.2 8.8 20.9 2.6E+03
30 3.5 132.6 8.9 20.8 2.5E+03
31 3.4 139.2 9.0 20.6 2.5E+03
32 3.2 146.1 9.1 20.5 2.4E+03
33 3.1 153.2 9.1 20.4 2.3E+03
34 3.0 160.5 9.2 20.2 2.3E+03
35 2.8 168.0 9.3 20.1 2.2E+03
36 2.7 175.8 9.4 19.9 2.2E+03
37 2.6 183.7 9.4 19.7 2.1E+03
38 2.5 191.9 9.5 19.5 2.0E+03
39 2.4 200.2 9.5 19.3 2.0E+03
40 2.3 208.7 9.6 19.1 1.9E+03
117
Nu h ∆T ƞo Qcal Fin
Number [-] [W/m2K] [K] [-] [W]
10 15.1 38.1 3.3 0.9954 37.9
11 14.6 38.6 3.5 0.9951 39.2
12 14.2 39.0 3.6 0.9948 40.4
13 13.7 39.4 3.7 0.9945 41.7
14 13.3 39.8 3.8 0.9943 42.9
15 12.9 40.2 3.9 0.9940 44.2
16 12.5 40.6 4.1 0.9938 45.3
17 12.1 40.9 4.2 0.9935 46.5
18 11.8 41.3 4.3 0.9933 47.6
19 11.4 41.6 4.4 0.9930 48.6
20 11.1 42.0 4.5 0.9928 49.7
21 10.8 42.3 4.6 0.9926 50.7
22 10.4 42.6 4.7 0.9924 51.6
23 10.1 42.8 4.8 0.9922 52.6
24 9.8 43.1 4.9 0.9920 53.5
25 9.5 43.3 5.0 0.9918 54.4
26 9.2 43.5 5.1 0.9917 55.2
27 9.0 43.7 5.2 0.9915 55.9
28 8.7 43.9 5.3 0.9914 56.6
29 8.4 44.0 5.4 0.9912 57.2
30 8.1 44.1 5.5 0.9911 57.9
31 7.9 44.2 5.6 0.9910 58.5
32 7.6 44.2 5.7 0.9909 58.9
33 7.3 44.2 5.7 0.9908 59.3
34 7.1 44.1 5.8 0.9907 59.6
35 6.8 44.0 5.9 0.9906 60.0
36 6.6 43.8 6.0 0.9906 60.2
37 6.3 43.5 6.0 0.9906 60.2
38 6.1 43.2 6.1 0.9905 60.3
39 5.8 42.8 6.1 0.9906 60.1
40 5.6 42.3 6.2 0.9906 60.0