Heat exchanger - Training course material

TRAINING MATERIAL
Heat Exchangers:
Design, Operation, Maintenance and Enhancement

2 Table of contents
Table of contents
1 Introduction 8
1.1 Programm outline . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8
1.2 Instructor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10
2 Classification of heat exchangers 12
2.1 Classification by construction . . . . . . . . . . . . . . . . . . . . . . . . . 14
2.1.1 Tubular heat exchanger . . . . . . . . . . . . . . . . . . . . . . . . 14
2.2 Double pipe heat exchanger . . . . . . . . . . . . . . . . . . . . . . . . . . 15
2.3 Spiral tube heat exchanger . . . . . . . . . . . . . . . . . . . . . . . . . . . 16
2.4 Shell and tube heat exchanger . . . . . . . . . . . . . . . . . . . . . . . . . 16
2.4.1 Fixed tubesheet . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17
2.4.2 U-tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 17
2.4.3 Floating head . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 18
2.5 Plate heat exchangers . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 19
2.5.1 Gasketed plate heat exchanger . . . . . . . . . . . . . . . . . . . . 20
2.5.2 Welded- and Brazed-Plate exchanger (W. PHE and BHE) . . . . . 22
2.5.3 Spiral Plate Exchanger (SPHE) . . . . . . . . . . . . . . . . . . . . 23
2.6 Extended surface . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26
2.6.1 Plate fin . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 26
2.6.2 Tube fin . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 27
3 Code and standards 28
3.1 TEMA Designations . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 28
3.2 Classification by construction STHE . . . . . . . . . . . . . . . . . . . . . 33
3.2.1 Fixed tube sheet . . . . . . . . . . . . . . . . . . . . . . . . . . . . 33
3.2.2 U-Tube Heat Exchanger . . . . . . . . . . . . . . . . . . . . . . . . 35
3.2.3 Floating Head Designs . . . . . . . . . . . . . . . . . . . . . . . . . 37
3.3 Shell Constructions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 41
3.4 Tube side construction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 42
3.4.1 Tube-Side Header: . . . . . . . . . . . . . . . . . . . . . . . . . . . 42
3.4.2 Tube-Side Passes . . . . . . . . . . . . . . . . . . . . . . . . . . . . 42
3.4.3 Tubes Type . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 43
3.4.4 Tube arrangement . . . . . . . . . . . . . . . . . . . . . . . . . . . 46
3.4.5 Tube side passes . . . . . . . . . . . . . . . . . . . . . . . . . . . . 47
3.5 Shell side construction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 47
3.5.1 Shell Sizes . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 47
3.5.2 Shell-Side Arrangements . . . . . . . . . . . . . . . . . . . . . . . . 48
3.6 Baffles and tube bundles . . . . . . . . . . . . . . . . . . . . . . . . . . . . 48
3.6.1 The tube bundle . . . . . . . . . . . . . . . . . . . . . . . . . . . . 48
Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : rabahss@hotamil.com

Table of contents 3
3.6.2 Baffle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 48
3.6.3 Vapor Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . . 51
3.6.4 Tube-Bundle Bypassing . . . . . . . . . . . . . . . . . . . . . . . . 51
3.6.5 Tie Rods and Spacers . . . . . . . . . . . . . . . . . . . . . . . . . . 52
3.6.6 Tubesheets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 52
4 Basic Design Equations of Heat Exchangers 55
4.1 LMTD-Method . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 55
4.1.1 Logarithmic mean temperature different . . . . . . . . . . . . . . . 56
4.1.2 Correction Factor . . . . . . . . . . . . . . . . . . . . . . . . . . . . 57
4.1.3 Overall heat transfer coefficient . . . . . . . . . . . . . . . . . . . . 59
4.1.4 Heat transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . . 61
4.1.5 Fouling factor (hid, hod) . . . . . . . . . . . . . . . . . . . . . . . . . 61
4.2 ε- NTU . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 61
4.3 Link between LMTD and NTU . . . . . . . . . . . . . . . . . . . . . . . . 64
4.4 The Theta Method . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 64
5 Thermal Design 66
5.1 Design Consideration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 66
5.1.1 Fluid Stream Allocations . . . . . . . . . . . . . . . . . . . . . . . . 66
5.1.2 Shell and tube velocity . . . . . . . . . . . . . . . . . . . . . . . . . 66
5.1.3 Stream temperature . . . . . . . . . . . . . . . . . . . . . . . . . . 67
5.1.4 Pressure drop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 67
5.1.5 Fluid physical properties . . . . . . . . . . . . . . . . . . . . . . . . 67
5.2 Design data . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 68
5.3 Tubeside design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 69
5.3.1 Heat-transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . . 69
5.3.2 Pressure drop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 70
5.4 Shell side design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 72
5.4.1 Shell configuration . . . . . . . . . . . . . . . . . . . . . . . . . . . 72
5.4.2 Tube layout patterns . . . . . . . . . . . . . . . . . . . . . . . . . . 73
5.4.3 Tube pitch . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 73
5.4.4 Baffling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 74
5.4.5 Equalize cross-flow and window velocities . . . . . . . . . . . . . . . 76
5.4.6 Shellside stream analysis (Flow pattern) . . . . . . . . . . . . . . . 76
5.4.7 Heat transfer coefficient and pressure drop . . . . . . . . . . . . . . 77
5.4.8 Heat transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . . 78
5.4.9 Pressure drop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 78
5.5 Design Algorithm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 79
6 Specification sheet 80

4 Table of contents
6.1 Information included . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80
6.2 Information not included . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80
6.3 Operation conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80
6.4 Bid evaluation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81
6.4.1 Factor to be consider . . . . . . . . . . . . . . . . . . . . . . . . . . 81
7 Storage, Installation, Operation and Maintenance 83
7.1 Storage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83
7.2 Installation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 85
7.2.1 Installation Planning . . . . . . . . . . . . . . . . . . . . . . . . . . 85
7.2.2 Installation at Jobsite . . . . . . . . . . . . . . . . . . . . . . . . . 86
7.3 Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 87
8 Heat exchanger tube side mainenance (Repair vs replacement 91
8.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 91
8.2 Repair vs. Replace - Factors To Consider . . . . . . . . . . . . . . . . . . . 92
8.3 Heat Exchanger maintenance Options . . . . . . . . . . . . . . . . . . . . . 93
8.4 Repair option . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94
8.4.1 Plug . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94
8.4.2 Sleeving . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95
8.4.3 Tube Expansion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 100
8.5 Replacement option . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 103
8.5.1 Retubing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 103
8.5.2 Rebundling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 104
8.5.3 Complete replacement (New unit) . . . . . . . . . . . . . . . . . . . 104
8.6 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 105
9 Troubleshooting 106
9.1 Heat exchangers’ problems . . . . . . . . . . . . . . . . . . . . . . . . . . . 106
9.2 Fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 106
9.2.1 Costs of fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 106
9.2.2 Facts about fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . 107
9.2.3 Types of Fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . . 107
9.2.4 Fouling Mechanisms . . . . . . . . . . . . . . . . . . . . . . . . . . 107
9.2.5 Conditions Inﬂuencing Fouling . . . . . . . . . . . . . . . . . . . . . 107
9.2.6 Fouling control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 108
9.2.7 Fouling cleaning methods . . . . . . . . . . . . . . . . . . . . . . . 108
9.3 Leakage/Rupture of the Heat Transfer Surface . . . . . . . . . . . . . . . . 109
9.3.1 Cost of leakage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 109
9.3.2 Cause of diﬀerential thermal expansion . . . . . . . . . . . . . . . . 109
9.4 Corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110

Table of contents 5
9.4.1 Corrosion effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110
9.4.2 Causes of corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . 110
9.4.3 Type of corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110
9.4.4 Stress corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110
9.4.5 Galvanic corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . 110
9.4.6 Pitting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111
9.4.7 Uniform or rust corrosion . . . . . . . . . . . . . . . . . . . . . . . 111
9.4.8 Crevice corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111
9.4.9 Materials of Construction . . . . . . . . . . . . . . . . . . . . . . . 112
9.4.10 Fabrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 112
9.5 Troubleshooting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 112
9.6 Past failure incidents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 113
9.6.1 Ethylene Oxide Redistillation Column Explosion: . . . . . . . . . . 113
9.6.2 Brittle Fracture of a Heat Exchanger . . . . . . . . . . . . . . . . . 113
9.6.3 Cold Box Explosion . . . . . . . . . . . . . . . . . . . . . . . . . . . 114
9.7 Failure scenarios and design solutions . . . . . . . . . . . . . . . . . . . . . 114
9.8 Discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 116
9.8.1 Use of Potential Design Solutions Table . . . . . . . . . . . . . . . . 116
9.8.2 Special Considerations . . . . . . . . . . . . . . . . . . . . . . . . . 117
9.9 Troubleshooting Examples . . . . . . . . . . . . . . . . . . . . . . . . . . . 118
9.9.1 Shell side temperature uncontrolled . . . . . . . . . . . . . . . . . . 118
9.9.2 Shell assumed banana-shape . . . . . . . . . . . . . . . . . . . . . . 118
9.9.3 Steam condenser performing below design capacity . . . . . . . . . 119
9.9.4 Steam heat exchanger flooded . . . . . . . . . . . . . . . . . . . . . 119
10 Unresolved problems in the heat exchangers design 120
10.1 Future trend . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 120
Bibliography 121
A Heat transfer coefficient 131
A.1 Single phase . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 131
A.1.1 Inside tube: Turbulent flow . . . . . . . . . . . . . . . . . . . . . . 131
A.1.2 Inside tube: Laminar flow . . . . . . . . . . . . . . . . . . . . . . . 131
A.1.3 Shell side . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 131
A.1.4 Plate heat exchanger . . . . . . . . . . . . . . . . . . . . . . . . . . 133
A.2 Condensation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 133
A.2.1 Condensation on vertical plate or outside vertical tube . . . . . . . 133
A.2.2 Condensation on external horizontal tube . . . . . . . . . . . . . . 133
A.2.3 Condensation on banks of horizontal tube . . . . . . . . . . . . . . 133
A.2.4 Condensation inside horizontal tube . . . . . . . . . . . . . . . . . . 134

6 Table of contents
A.3 Two phase flow: Pure fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . 134
A.3.1 Steiner [140] correlation . . . . . . . . . . . . . . . . . . . . . . . . 134
A.3.2 Kattan et al. [77] correlation . . . . . . . . . . . . . . . . . . . . . . 137
A.3.3 Kandlikar [70] correlation . . . . . . . . . . . . . . . . . . . . . . . 138
A.3.4 Chen [19] correlation . . . . . . . . . . . . . . . . . . . . . . . . . . 139
A.3.5 Gungor and Winterton [52] correlation . . . . . . . . . . . . . . . . 140
A.3.6 Shah [130] correlation . . . . . . . . . . . . . . . . . . . . . . . . . . 140
A.3.7 Schrock and Grossman [129] correlation . . . . . . . . . . . . . . . . 141
A.3.8 Dembi et al. [30] correlation . . . . . . . . . . . . . . . . . . . . . . 141
A.3.9 Klimenko [84] correlation . . . . . . . . . . . . . . . . . . . . . . . . 141
A.3.10 Jung et al. [64] correlation . . . . . . . . . . . . . . . . . . . . . . . 142
A.4 Two phase flow: Mixture . . . . . . . . . . . . . . . . . . . . . . . . . . . . 142
A.4.1 Steiner [140] correlation . . . . . . . . . . . . . . . . . . . . . . . . 142
A.4.2 Kandlikar [71] correlation . . . . . . . . . . . . . . . . . . . . . . . 143
A.4.3 Bennett and Chen [8] correlation . . . . . . . . . . . . . . . . . . . 143
A.4.4 Palen [111] correlation . . . . . . . . . . . . . . . . . . . . . . . . . 143
A.4.5 Jung et al. [64] correlation . . . . . . . . . . . . . . . . . . . . . . . 144
B Pressure drop 145
B.1 Single phase . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 145
B.2 Two phase . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 145
B.2.1 Friedel [42] model . . . . . . . . . . . . . . . . . . . . . . . . . . . 147
B.2.2 Lockhart and Martinelli [91] model . . . . . . . . . . . . . . . . . . 147
B.2.3 Chisholm [22] model . . . . . . . . . . . . . . . . . . . . . . . . . . 148
C Physical properties 149
C.1 Physical properties: Pure fluid . . . . . . . . . . . . . . . . . . . . . . . . . 149
C.1.1 Specific heat . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149
C.1.2 Vapor pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149
C.1.3 Liquid viscosity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149
C.1.4 Vapor dynamic viscosity VDI-Wärmeatlas [157] . . . . . . . . . . . 149
C.1.5 Dynamic viscosity of Fenghour et al. [40] . . . . . . . . . . . . . . . 151
C.1.6 Surface tension . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 152
C.1.7 Thermal conductivity for liquids . . . . . . . . . . . . . . . . . . . . 152
C.1.8 Thermal conductivity for gases . . . . . . . . . . . . . . . . . . . . 152
C.1.9 Specific enthalpy . . . . . . . . . . . . . . . . . . . . . . . . . . . . 153
C.2 Physical properties: Mixture . . . . . . . . . . . . . . . . . . . . . . . . . . 153
C.2.1 Liquid dynamic viscosity of mixtures . . . . . . . . . . . . . . . . . 153
C.2.2 Vapor dynamic viscosity of mixtures . . . . . . . . . . . . . . . . . 153
C.2.3 Liquid thermal conductivity of mixtures . . . . . . . . . . . . . . . 154

Table of contents 7
C.2.4 Vapor thermal conductivity of mixtures . . . . . . . . . . . . . . . . 154
C.2.5 Surface tension of mixtures . . . . . . . . . . . . . . . . . . . . . . 155
C.3 Software packages . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 155

8 1 Introduction
1 Introduction
Heat exchanger is an important and expensive item of equipment that is used almost in
every industry (oil and petrochemical, sugar, food, pharmaceutical and power industry).
A better understanding of the basic principles of heat transfer and fluid flow and their
application to the design and operation of heat exchangers that you gain from this course
will enable you to improve their efficiency and extend their life. You understand how to use
the applicable API, TEMA and ASME recommended practices, standards and codes for
heat exchangers. This will enable you to communicate with the designers, manufacturers
and bidders of heat exchangers. You will understand how to avoid fouling, corrosion and
failure and leak problems by your design. You will also be able to survey and troubleshoot
heat exchangers and assist in performing inspection, cleaning, and maintenance. You will
be exposed to recent development and future trend in heat exchangers.
The course includes worked examples to reinforce the key learning as well as a demon-
stration of mechanical design and challenging problems encountered in the operation of
heat exchangers.
Objectives
• To learn the classification, code and standards (API, TEMA,...) and selection pro-
cedure for heat exchangers.
• To review the thermal and mechanical design of heat exchangers.
• To learn the installation, operation and maintenance procedure for heat exchanger.
• To acquire information that will enable decisions to be made on the repair and
refurbishment of aging equipment as well as repair vs. replacement options.
• To learn techniques of failure elimination and appropriate maintenance and trou-
bleshooting procedures.
• To delineate the factors that lead to overall economically advantageous decisions.
Who should attend: Project engineers, process engineers and plant engineers in the oil,
chemical, sugar, power, and other industries who requires a wider and deeper appreciation
of heat exchangers design, performance and operation. The detailed review of thermal
and mechanical design is particularly useful to plant and maintenance engineers as well
as to those generally knowledgeable in the subject, but who require a refresher or up-
date. Codes and standards are useful for project engineer to help him communicate with
manufacturers, designers and bidders of heat exchangers. Troubleshooting procedures are
important for process engineers. Participants will be taken through an intensive primer
of heat transfer principles as applicable to heat exchangers.
1.1 Programm outline
1. DAY I: HEAT EXCHANGERS CLASSIFICATION APPLICATION, CODE
AND STANDARDS
• Classification according to construction (tubular, plate, finned, enhanced)
• Classification according to service (cooler, heater, condenser, reboiler, etc..)
• Construction, applications, range and limitations and sizes
• Code and standards (TEMA, API,...)
• TEMA nomenclature: rear end head types, shell types, font end types
• TEMA standards: shell size, tube size, baffle, selection of materials, component
design, nozzle loadings, supports, lifting features, high pressure, low tempera-
ture, specials designs

1.1 Programm outline 9
2. DAY II HEAT TRANSFER FUNDAMENTALS AND THERMAL DE-
SIGN
• Heat transfer mechanisms: conduction and convection as related to heat ex-
changers
• Temperature difference in heat exchanger:
– LMTD Method
– ε-NTU Method
– θ-Method
• Overall heat transfer coefficient
• Heat transfer coefficient and pressure drop for single phase and multiphase
(evaporation and condensation)
• Resistances to fouling
• Illustration examples using the software CHEMCAD
3. DAY III MECHANICAL DESIGN OF HE
• Mechanical design: shells, channels and heads, tubesheets, bundles, tubes-
tubesheet attachment
• Design strategy, design algorithm
• Heat exchanger:
– Selection procedure
– Specification sheet
– Bid evaluation
• Worked example (USING CHEMCAD)
4. DAY IV Storage, Installation, Operation, Maintenance
• Storage
• Installation procedure
• Operation
• start up
• shut down
• Maintenance
• Cleaning
• Repair
– Plug
– Sleeving
– Expansion
• Replacement
– Retubing
– Rebundling
– Replacement (new unit)
5. DAY V Troubleshooting
• Heat exchangers’ problem

10 1 Introduction
– Fouling: causes, mechanisms, design considerations and exchanger selec-
tion, remedies, cleaning
– Leakage: Location (tube sheet, tube failure), causes (differential thermal
expansion, flow-induced vibration),
– Corrosion: Type, causes, material of construction, fabrication
– Vibration: causes (velocity), design procedure to avoid vibration including
baffle selection, rod baffles, impingement baffles
• Past incidents failure.
• Examples of common problems encountered in heat exchangers (low rate, un-
controlled outlet temperature, failure of tubes near the inlet nozzles)
Achieve the learning outcomes to:
Understand the principles of heat transfer and fluid flow, application of industry prac-
tices and a substantial amount of supporting data needed for design, performance and
operation of modern heat exchangers.
Gain insight not only into shell and tube heat exchangers but also heat transfer funda-
mentals as applied to heat exchangers, the types of heat exchangers and their application,
and recent advance in heat exchanger technologies
Become familiar with the practical aspects and receive tips on shell and tube heat
exchanger thermal design and rating: mechanical design and rating using the applicable
API, TEMA and ASME recommended practices, standards and codes, troubleshooting,
and performance improvement and enhancement
Avoid future problems by gaining insight into vibration forcing mechanisms
Enhance your awareness of causes of failure and learn practical ways for determining
and correcting them
Daily Schedule: 8:00 Registration and Coffee (1st day only) 8:30 Session begins 4:30
Adjournment
There will be a forty-minute lunch break each day in addition to refreshment and net-
working break of 20 minutes during each morning and afternoon session.
1.2 Instructor
Faculty: Ali. Rabah, BSc. MSc., PhD., MSES., Assistant professor, De-
partment of Chemical Engineering University of Khartoum
Dr. Rabah holds a BSc. degree (Chemical Engineering) from the University of Khartoum,
MSc. degree from university of Nairobi, Kenya, and PhD. degree from University of
Hannover, Germany. He has a wide professional experience in teaching heat and mass
transfer and engineering thermodynamics to BSc and MSc Chemical, Mechanical and
Petroleum Engineering students.
Dr. Rabah is a consultant engineer to a number of chemical industries and factories.
He has developed and delivered numerous designs of heat exchangers, evaporators and
boilers. He designed, for example, a 5 ton/hr (10 bar) fired tube boiler. His design is
under fabrication.
Dr. Rabah has designed and manufactured double pipe heat exchangers for education
proposes to a number of chemical engineering departments country-wide e.g. University
of Nileen.
Dr. Rabah assumed engineering design positions with responsibilities covering design,
construction and inspection of heat transfer equipments. The design projects are spon-
sored by the federal ministry of research and technology and the University of Khartoum
consultancy cooperation.

1.2 Instructor 11
Dr. Rabah is a member of the Sudan Engineering Society (SES) and serving as a member
of editorial board of SES Journal. He is a reviewer to a number of world wide soft-
ware packages for chemical engineering simulations and the prediction of thermodynamic
properties.
Dr. Rabah has a number of publications in ﬁeld of heat transfer and thermodynamics.

12 2 Classification of heat exchangers
2 Classification of heat exchangers
The word exchanger really applies to all types of equipment in which heat is exchanged but
is often used specially to denote equipment in which heat is exchanged between two process
streams. Exchangers in which a process fluid is heated or cooled by a plant service stream
are referred to as heatsers and coolers. If the process stream is vaporized the exchanger is
called a vaporizer if the the stream is essentially completely vaporized: called a reboiled
if associated with a distillation column: and evaporator if used to concentrate a solution.
If the process fluid is condensed the exchanger is called a condenser. The term fired
exchanger is used for exchangers heated by combustion gases, such as boiler. In heat
exchanger the heat transfer between the fluid takes place through a separating wall. The
wall may a solid wall or interface. Heat exchangers are used in
• Oil and petrochemical Industry (upstream and down stream)
• Sugar industry
• Power generation industry
• Air-cooling and refrigeration industry
These heat exchanger may be classified according to:
• Transfer process
1. Direct contact
2. indirect contact
(a) Direct transfer type
(b) Storage type
(c) Fluidized bed
• Surface compactness
1. Compact (surface area density ≥ 700m2
/m3
)
2. non-compact (surface area density < 700m2
/m3
)
• Construction
1. Tubular
(a) Double pipe
(b) Shell and tube
(c) Spiral tube
2. Plate
(a) Gasketed
(b) Spiral plate
(c) Welded plate
3. Extended surface
(a) Plate fin
(b) Tube fin
4. Regenerative
(a) Rotory
i. Disc-type

13
ii. Drum-type
(b) Fixed-matrix
• Flow arrangement
1. Single pass
(a) Parallel flow
(b) Counter flow
(c) Cross flow
2. Multipass
(a) Extended surface H.E.
i. Cross counter flow
ii. Cross parallel flow
(b) Shell and tube H.E.
i. Parallel counter flow (Shell and fluid mixed, M shell pass, N Tube pass)
ii. Split flow
iii. Divided flow
(c) Plate H.E. (N-parallel plate multipass)
• Number of fluids
1. Two-fluid
2. Three fluid
3. N-fluid (N > 3)
• Transfer mechanisms
1. Single phase convection on both sides
2. Single phase convection on one side, two-phase convection on the other side
3. Two-phase convection on both sides
4. Combined convection and radiative heat transfer
• Classification based on service: Basically, a service may be single phase (such as the
cooling or heating of a liquid or gas) or two-phase (such as condensing or vaporizing).
Since there are two sides to an STHE, this can lead to several combinations of ser-
vices. Broadly, services can be classified as follows: single-phase (both shellside and
tubeside); condensing (one side condensing and the other single-phase); vaporizing
(one side vaporizing and the other side single-phase); and condensing/vaporizing
(one side condensing and the other side vaporizing). The following nomenclature is
usually used:
– Heat exchanger: both sides singlephase and process streams (that is, not a
utility).
– Cooler: one stream a process fluid and the other cooling water or air. Dirty
water can be used as the cooling medium. The top of the cooler is open to the
atmosphere for access to tubes. These can be cleaned without shutting down
the cooler by removing the distributors one at a time and scrubbing the tubes.
– Heater: one stream a process fluid and the other a hot utility, such as steam
or hot oil.
– Condenser: one stream a condensing vapor and the other cooling water or air.

– Chiller: one stream a process fluid being condensed at sub-atmospheric tem-
peratures and the other a boiling refrigerant or process stream. By cooling the
falling film to its freezing point, these exchangers convert a variety of chemicals
to the solid phase. The most common application is the production of sized ice
and paradichlorobenzene. Selective freezing is used for isolating isomers. By
melting the solid material and refreezing in several stages, a higher degree of
purity of product can be obtained.
– Reboiler: one stream a bottoms stream from a distillation column and the
other a hot utility (steam or hot oil) or a process stream.
– Evaporators:These are used extensively for the concentration of ammonium
nitrate, urea, and other chemicals sensitive to heat when minimum contact
time is desirable. Air is sometimes introduced in the tubes to lower the partial
pressure of liquids whose boiling points are high. These evaporators are built
for pressure or vacuum and with top or bottom vapor removal.
– Absorbers: These have a two-phase flow system. The absorbing medium is
put in film flow during its fall downward on the tubes as it is cooled by a cooling
medium outside the tubes. The film absorbs the gas which is introduced into
the tubes. This operation can be cocurrent or countercurrent.
– Falling-Film Exchangers: Falling-film shell-and-tube heat exchangers have
been developed for a wide variety of services and are described by Sack [Chem.
Eng. Prog., 63, 55 (July 1967)]. The fluid enters at the top of the vertical
tubes. Distributors or slotted tubes put the liquid in film flow in the inside
surface of the tubes, and the film adheres to the tube surface while falling
to the bottom of the tubes. The film can be cooled, heated, evaporated, or
frozen by means of the proper heat-transfer medium outside the tubes. Tube
distributors have been developed for a wide range of applications. Fixed tube
sheets, with or without expansion joints, and outside-packed-head designs are
used. Principal advantages are high rate of heat transfer, no internal pressure
drop, short time of contact (very important for heat-sensitive materials), easy
accessibility to tubes for cleaning, and, in some cases, prevention of leakage
from one side to another. These falling-film exchangers are used in various
services as described in the following paragraphs.
Among these classifications the classification by construction is the most widely used one.
2.1 Classification by construction
The principal types of heat exchanger are listed again as
1. Tubular exchanger
2. Plate exchanger
3. Extended surface
4. Regenerative
2.1.1 Tubular heat exchanger
Tubular heat exchanger are generally built of circular tubes. Tubular heat exchanger is
further classified into:
• Double pipe heat exchanger
• Spiral tube heat exchanger
• Shell and tube heat exchanger

2.2 Double pipe heat exchanger 15
2.2 Double pipe heat exchanger
This is usually consists of concentric pipes. One fluid flow in the inner pipe and the other
fluid flow in the annulus between pipes. The two fluid may flow concurrent (parallel) or
in counter current flow configuration; hence the heat exchanger are classified as:
• counter current double pipe heat exchanger (see Fig. 4.1and Fig. 2.2)and
• cocurrent double pipe heat exchanger
Figure 2.1. Double pipe heat exchanger. Courtesy of Perry, Chemical engineering hand book
Flowmeter
Bypass
pump
Tee 2"x1/2"
Union 2"
Galv. pipe 2"
Cu pipe 3/4"
Tee 3/4"x1/2"
Elbew 3/4"
Flanged Gland 2"
Part B
Double Pipe Heat Exchanger
Scale: None Sheet No.1 Date: 08.12.2003
Designed by: Dr.-Ing. Ali A. Rabah
Part A
Specification Sheet
Item Qty Item Qty
Tee 2"x3/4" 6 Tee 3/4"x1/2" 14
Union 2" 6 Cu Bush 1/2" 8
Valve 3/4" 4 Elbew 3/4" 10
Galv. pipe 2"x3ft 3 Cu pipe 3/4"x4ft 3
Galv. pipe 3/4"x1ft Selector
(Threaded) 24 (20 Channel) 1
Cu Flange 2" 8 Flow meter 3/4" 2
Pump 0-40 l/min 2 Union 3/4" 30
Amplifier 1 Microvoltmeter 1
Thermocouples Elbew 1/2" 4
(NiCr-Ni) 10 Union 1/2" 8
Valve3/4"
Galv. pipe
Threaded 3/4"
Bypass
Figure 2.2. Double pipe heat exchanger (Counter current)
Double pipe heat exchanger is perhaps the simplest of all heat exchanger types. The
advantages of this type are:

i Easily by disassembly, no cleaning problem
ii Suitable for high pressure fluid, (the pressure containment in the small diameter pipe
or tubing is a less costly method compared to a large diameter shell.)
Limitation: The double pipe heat exchanger is generally used for the application where
the total heat transfer surface area required is less than or equal to 20 m2
(215 ft2
) because
it is expensive on a cost per square meter (foot) basis.
2.3 Spiral tube heat exchanger
Spiral tube heat exchanger consists of one or more spirally wound coils fitted in a shell
(Fig. 2.3). Heat transfer associated with spiral tube is higher than than that for a straight
tube . In addition, considerable amount of surface area can be accommodated in a given
space by spiralling. Thermal expansion is no problem but cleaning is almost impossible.
Figure 2.3. Spiral tube heat exchanger. Courtesy of The German Atlas
2.4 Shell and tube heat exchanger
Shell and tube heat exchanger is built of round tubes mounted in a cylindrical shell with
the tube axis parallel to that of the shell. One fluid flow inside the tube, the other flow
across and along the tubes. The major components of the shell and tube heat exchanger
are tube bundle, shell, front end head, rear end head, baffles and tube sheets (Fig.2.4).
Figure 2.4. Shell and tube heat exchanger

2.4 Shell and tube heat exchanger 17
The shell and tube heat exchanger is further divided into three catogaries as
1. Fixed tube sheet
2. U tube
3. Floating head
2.4.1 Fixed tubesheet
A fixed-tubesheet heat exchanger (Figure 2.5) has straight tubes that are secured at both
ends to tubesheets welded to the shell. The construction may have removable channel
covers , bonnet-type channel covers , or integral tubesheets. The principal advantage of
the fixedtubesheet construction is its low cost because of its simple construction. In fact,
the fixed tubesheet is the least expensive construction type, as long as no expansion joint
is required.
Figure 2.5. Fixed-tubesheet heat exchanger.
Other advantages are that the tubes can be cleaned mechanically after removal of the
channel cover or bonnet, and that leakage of the shellside fluid is minimized since there
are no flanged joints.
A disadvantage of this design is that since the bundle is fixed to the shell and cannot be
removed, the outsides of the tubes cannot be cleaned mechanically. Thus, its application
is limited to clean services on the shellside. However, if a satisfactory chemical clean-
ing program can be employed, fixed-tubesheet construction may be selected for fouling
services on the shellside.
In the event of a large differential temperature between the tubes and the shell, the
tubesheets will be unable to absorb the differential stress, thereby making it necessary to
incorporate an expansion joint. This takes away the advantage of low cost to a significant
extent.
2.4.2 U-tube
As the name implies, the tubes of a U-tube heat exchanger (Figure 2.6) are bent in
the shape of a U. There is only one tubesheet in a Utube heat exchanger. However,
the lower cost for the single tubesheet is offset by the additional costs incurred for the
bending of the tubes and the somewhat larger shell diameter (due to the minimum U-bend
radius), making the cost of a U-tube heat exchanger comparable to that of a fixedtubesheet
exchanger.

The advantage of a U-tube heat exchanger is that because one end is free, the bundle
can expand or contract in response to stress differentials. In addition, the outsides of the
tubes can be cleaned, as the tube bundle can be removed.
The disadvantage of the U-tube construction is that the insides of the tubes cannot be
cleaned effectively, since the U-bends would require flexible- end drill shafts for cleaning.
Thus, U-tube heat exchangers should not be used for services with a dirty fluid inside
tubes.
Figure 2.6. U-tube heat exchanger.
2.4.3 Floating head
The floating-head heat exchanger is the most versatile type of STHE, and also the costliest.
In this design, one tubesheet is fixed relative to the shell, and the other is free to ”float”
within the shell. This permits free expansion of the tube bundle, as well as cleaning
of both the insides and outsides of the tubes. Thus, floating-head SHTEs can be used
for services where both the shellside and the tubeside fluids are dirty-making this the
standard construction type used in dirty services, such as in petroleum refineries.
There are various types of floating- head construction. The two most common are the
pull-through with backing device and pullthrough without backing service designs. The
design (Figure 2.7) with backing service is the most common configuration in the chemical
process industries (CPI). The floating-head cover is secured against the floating tubesheet
by bolting it to an ingenious split backing ring. This floating-head closure is located
beyond the end of the shell and contained by a shell cover of a larger diameter. To
dismantle the heat exchanger, the shell cover is removed first, then the split backing ring,
and then the floating-head cover, after which the tube bundle can be removed from the
stationary end.

2.5 Plate heat exchangers 19
Figure 2.7. Floating head with packing service.
In the design without packing service construction (Figure 2.8), the entire tube bundle,
including the floating-head assembly, can be removed from the stationary end, since the
shell diameter is larger than the floating-head flange. The floatinghead cover is bolted
directly to the floating tubesheet so that a split backing ring is not required. The advan-
tage of this construction is that the tube bundle may be removed from the shell without
removing either the shell or the floatinghead cover, thus reducing maintenance time. This
design is particularly suited to kettle reboilers having a dirty heating medium where U-
tubes cannot be employed. Due to the enlarged shell, this construction has the highest
cost of all exchanger types.
Figure 2.8. Floating head without packing service.
2.5 Plate heat exchangers
These exchangers are generally built of thin plates. The plate are either smooth or have
some form of corrugations and they are either flat or wound in exchanger. Generally

theses exchanger cannot accomodate high pressure/temperature differential relative the
tubular exchanger. This type of exchanger is further classified as:
• Gasketed plate
• Fixed plate
• Spiral plate
2.5.1 Gasketed plate heat exchanger
Gasketed plate heat exchanger (see Fig. 2.9) consists of a series of corrugated alloy
material channel plates, bounded by elastomeric gaskets are hung off and guided by lon-
gitudinal carrying bars, then compressed by large-diameter tightening bolts between two
pressure retaining frame plates (cover plates).
Figure 2.9. Plate heat exchanger
The frame and channel plates have portholes which allow the process fluids to enter alter-
nating flow passages (the space between two adjacent-channel plates) Fig.2.10. Gaskets
around the periphery of the channel plate prevent leakage to the atmosphere and also pre-
vent process fluids from coming in contact with the frame plates. No inter fluid leakage
is possible in the port area due to a dual-gasket seal. Fig.2.11 shows the plate profiles.
Expansion of the initial unit is easily performed in the field without special considerations.
The original frame length typically has an additional capacity of 15-20 percent more
channel plates (i.e. surface area). In fact, if a known future capacity is available during
fabrication stages, a longer carrying bar could be installed, and later, increasing the
surface area would be easily handled. When the expansion is needed, simply untighten
the carrying bolts, pull back the frame plate, add the additional channel plates, and
tighten the frame plate.

Figure 2.10. Plate heat exchanger flow configuration
Applications: Most PHE applications are liquid-liquid services but there are numerous
steam heater and evaporator uses from their heritage in the food industry. Industrial users
typically have chevron style channel plates while some food applications are washboard
style.
Fine particulate slurries in concentrations up to 70 percent by weight are possible with
standard channel spacings. Wide-gap units are used with larger particle sizes. Typical
particle size should not exceed 75 percent of the single plate (not total channel) gap.
Close temperature approaches and tight temperature control possible with PHE’s and the
ability to sanitize the entire heat transfer surface easily were a major benefit in the food
and pharmaceutical industry.
Advantages: -
• Easily assembled and dismantled
• Easily cleaned both chemically and mechanically
• Flexible (the heat transfer can be changed as required)
• Can be used for multiple service as required
• Leak is immediately deteced since all plates are vented to the atmosphere, and the
fluid split on the floor rather than mixing with other fluid
• Heat transfer coefficient is larger and hence small heat transfer area is required than
STHE
• The space required is less than that for STHE for the same duty
• Less fouling due to high turbulent flow

Figure 2.11. Plate and frame of a plate heat exchanger
• Very close temperature approach can be obtained
• low hold up volume
• LMTD is fully utilized
• More economical when material cost are high
Disadvantages: -
• Low pressure <30 bar (plate deformation)
• Working temperature of < (500 F) [250 o
C] (maximum gasket temperature) see
table 2.1.
Table 2.1. Plate Heat Exchanger Gasket Materials
Material Common name Temperature limit (F)
Styrene-Butadiene Buna-S 185
Neoprene Neoprene 250
Acrylonitrile- Butadiene Buna-N 275
Ethylene/Propylene EPDM 300
Fluorocarbon Viton 300
Resin-Cured Butyl Resin-Cured Butyl 300
Compressed Asbestos Compressed Asbestos 500
2.5.2 Welded- and Brazed-Plate exchanger (W. PHE and BHE)
To overcome the gasket limitations, PHE manufacturers have developed welded-plate
exchangers. There are numerous approaches to this solution: weld plate pairs together
with the other ﬂuid-side conventionally gasketed, weld up both sides but use a horizonal

stacking of plates method of assembly, entirely braze the plates together with copper or
nickel brazing, diffusion bond then pressure form plates and bond etched, passage plates
Fig. 2.12 and Fig. 2.13.
Typical applications include district heating where the low cost and minimal maintenance
have made this type of heat exchanger especially attractive.
Figure 2.12. Welded or blazed plate heat exchanger
Figure 2.13. Fin-Plate heat exchanger
Most methods of welded-plate manufacturing do not allow for inspection of the heat-
transfer surface, mechanical cleaning of that surface, and have limited ability to repair
or plug off damage channels. Consider these limitations when the fluid is heavily fouling,
has solids, or in general the repair or plugging ability for severe services.
2.5.3 Spiral Plate Exchanger (SPHE)
The spiral-plate heat exchanger (SHE) may be one exchanger selected primarily on its
virtues and not on its initial cost. SPHEs offer high reliability and on-line performance in
many severely fouling services such as slurries. The SHE is formed by rolling two strips

of plate, with welded-on spacer studs, upon each other into clock-spring shape Fig.2.14
and Fig.2.15. This forms two passages. Passages are sealed off on one end of the SHE by
welding a bar to the plates; hot and cold fluid passages are sealed off on opposite ends of
the SHE. A single rectangular flow passage is now formed for each fluid, producing very
high shear rates compared to tubular designs. Removable covers are provided on each
end to access and clean the entire heat transfer surface.
Figure 2.14. Spiral Plate heat exchanger
Pure countercurrent flow is achieved and LMTD correction factor is essentially = 1.0.
Since there are no dead spaces in a SHE, the helical flow pattern combines to entrain
any solids and create high turbulence creating a self-cleaning flow passage. There are
no thermal-expansion problems in spirals. Since the center of the unit is not fixed, it
can torque to relieve stress. The SHE can be expensive when only one fluid requires a

high alloy material. Since the heat-transfer plate contacts both fluids, it is required to be
fabricated out of the higher alloy. SHEs can be fabricated out of any material that can be
cold-worked and welded. The channel spacings can be different on each side to match the
flow rates and pressure drops of the process design. The spacer studs are also adjusted in
their pitch to match the fluid characteristics. As the coiled plate spirals outward, the plate
thickness increases from a minimum of 2 mm to a maximum (as required by pressure)
up to 10 mm. This means relatively thick material separates the two fluids compared to
tubing of conventional exchangers.
a) Spiral flow in both channels b) Flow are both spiral and axial
Figure 2.15. Spiral Plate heat exchanger
Applications: The most common applications that fit SHE are slurries. The rectan-
gular channel provides high shear and turbulence to sweep the surface clear of blockage
and causes no distribution problems associated with other exchanger types. A localized
restriction causes an increase in local velocity which aids in keeping the unit free flowing.
Only fibers that are long and stringy cause SHE to have a blockage it cannot clear itself.
As an additional antifoulant measure, SHEs have been coated with a phenolic lining. This
provides some degree of corrosion protection as well, but this is not guaranteed due to
pinholes in the lining process.
There are three types of SHE to fit different applications:
• Type I is the spiral-spiral flow pattern (Fig. 2.15a). It is used for all heating and
cooling services and can accommodate temperature crosses such as lean/rich services
in one unit. The removable covers on each end allow access to one side at a time to
perform maintenance on that fluid side. Never remove a cover with one side under
pressure as the unit will telescope out like a collapsible cup.

• Type II units are the condenser and reboiler designs (Fig. 2.15b). One side is spiral
flow and the other side is in cross flow. These SHEs provide very stable designs
for vacuum condensing and reboiling services. A SHE can be fitted with special
mounting connections for reflux-type ventcondenser applications. The vertically
mounted SHE directly attaches on the column or tank.
• Type III units are a combination of the Type I and Type II where part is in spiral
flow and part is in cross flow. This SHE can condense and subcool in a single
unit. The unique channel arrangement has been used to provide on-line cleaning,
by switching fluid sides to clean the fouling (caused by the fluid that previously
flowed there) off the surface. Phosphoric acid coolers use pond water for cooling
and both sides foul; water, as you expect, and phosphoric acid deposit crystals. By
reversing the flow sides, the water dissolves the acid crystals and the acid clears up
the organic fouling. SHEs are also used as oleum coolers, sludge coolers/ heaters,
slop oil heaters, and in other services where multiple flow- passage designs have not
performed well.
2.6 Extended surface
The tubular and plate exchangers described previously are all prime surface heat exchang-
ers. The design thermal effectiveness is usually 60 % and below and the heat transfer area
density is usually less than 300 m2
m3
. In many application an effectiveness of up to 90
% is essential and the box volume and mass are limited so that a much more compact
surface is mandated. Usually either a gas or a liquid having a low heat transfer coefficient
is the fluid on one or both sides. This results in a large heat transfer area requirements.
for low density fluid (gases), pressure drop constraints tend to require a large flow area.
so a question arises how can we increase both the surface area and flow area together in
a reasonably shaped configuration.
The surface area may be increased by the fins. The flow area is increased by the use of
thin gauge material and sizing the core property. There are two most common types of
extended surface heat exchangers. These are
• Plate-fin
• Tube-fin
2.6.1 Plate fin
Plate -fin heat exchanger has fins or spacers sandwiched between parallel plates (refereed
to as parting plates or parting sheets) or formed tubes as shown in fig. 2.16(left). While
the plates separate the two fluid streams, the fins form the individual flow passages. Fins
are used on both sides in a gas-gas heat exchanger. In gas-liquid applications fins are
used in the gas side.
Figure 2.17. Finned tube heat exchanger

2.6 Extended surface 27
Figure 2.16. Examples of extended surfaces on one or both sides. Plate fins on both sides
(left) and Tubes and plate fins (right).
2.6.2 Tube fin
In tube fin heat exchanger, tubes of round, rectangular, or elliptical shape are generally
used. Fins are generally used on the outside and also used inside the tubes in some
applications. they are attached to the tube by tight mechanical fit, tension wound, gluing,
soldering, brazing, welding or extrusion. Tube fin exchanger is shown in Fig. 2.16(right)
and Fig.2.17

28 3 Code and standards
3 Code and standards
The objective of codes and standards are best described by ASME: The objectives of
code rules and standards (apart from fixing dimensional values) is to achieve minimum
requirements for safe construction, in other words, to provide public protection by defining
those materials, design, fabrication and inspection requirements; whose omission may
radically increase operating hazards.... Experience with code rules has demonstrated that
the probability of disastrous failure can be reduced to the extremely low level necessary to
protect life and property by suitable minimum requirements and safety factors. Obviously,
it is impossible for general rules to anticipate other than conventional service,.... Suitable
precautions are therefore entirely the responsibility of the design engineer guided by the
needs and specifications of the user.
Over years a number of standardization bodies have been developed by individual country,
manufacturers and designers to lay down nomenclatures for the size and type of shell and
tube heat exchangers. These include among other
• TEMA standards (Tubular Exchanger Manufacturer Association., 1998)[147]
• HEI standards (Heat Exchanger Institute, 1980),
• API (American Petroleum Institute).
• Other national standards include the German (DIN), Japan, India, to mention a
few.
In this work, being most widely used one, the TEMA standard is presented.
3.1 TEMA Designations
In order to understand the design and operation of the shell and tube heat exchanger, it
is important to know the nomenclature and terminology used to describe them and the
various parts that go to their construction. Only then we can understand the design and
reports given by the researchers, designers, manufacturer and users.
It is essential for the designer to have a good working knowledge of the mechanical features
of STHEs and how they influence thermal design. The principal components of an STHE
are:
• shell;
• shell cover;
• tubes;
• channel;
• channel cover;
• tubesheet;
• baffles; and
• nozzles.
Other components include tie-rods and spacers, pass partition plates, impingement plate,
longitudinal baffle, sealing strips, supports, and foundation. Table 3.1 shows the nomen-
clature used for different parts of shell and tube exchanger in accordance with TEMA
standards; the numbers refer to the feature shown in Fig. 3.2 to Fig. 3.8.

3.1 TEMA Designations 29
Table 3.1. TEAM notations
Index Notation Index Notation
1 stationary head- channel 20 slip on backing flange
2 stationary head- bonnet 21 floating head cover-external
3 stationary head flange-chennel or bonnet 22 floating tube sheet skirt
4 channel cover 23 packing box
5 stationary head - nozzle 24 packing
6 stationary tube sheet 25 packing gland
7 tubes 26 latern ring
8 shell 27 tie rods and spacers
9 shell cover 28 traverse baffle or support plate
10 shell flange-stationary head end 29 impingement plate
11 shell flange-rear head end 30 longitudinal baffle
12 shell nozzle 31 pass partition
13 shell cover flange 32 vent connection
14 expansion joint 33 drain connection
15 floating tube sheet 34 instrument connection
16 floating head cover 35 support saddle
17 floating head flange 36 lifting lug
18 floating head backing device 37 support bracket
19 split shear ring 38 weir
39 liquid level connection
Because of the number of variations in mechanical designs for front and rear heads and
shells, and for commercial reasons, TEMA has divided STHE into main three components:
front head, shell and rear head. Fig. 3.1 illustrates TEMA nomenclature for the various
construction possibilities. TEMA has classified the front head channel and bonnet types as
given the letters (A,B,C,N,D) and the shell is classified according to the nozzles locations
for the inlet and outlet. There are type of shell configuration ( E,F,G,H,J,K,X). Similarly
the rear head is classified ( M,N,P,S,T,U,W).
Exchangers are described by the letter codes of the three sections. The first letter stands
for the front head, the second letter for the shell type and the third letter for the rear head
type. For example a BFL exchanger has a bonnet cover, two-shell pass with longitudinal
baffles and a fixed tube sheet rear head.
In addition to these the size of the exchanger is required to be identified with the notation.
The size is identified by the shell inside diameter (nominal) and tube length (both are
rounded to the nearest integer in inch or mm). Demonstration examples are shown below:
• Type AES size 23-192 in (590-4880): This exchanger has a removable channel
cover (A), single pass shell (E) and Split ring floating front head (S) it has , 23 in
(590 mm) inside diameter with tubes of 16 ft (4880 mm) long.
• Type BGU Size 19-84 (480-2130)This exchanger has a bonnet-type stationary
front head (B), split flow shell (G) and U-tube bundle rear head(U) with 19 in (480)
inside diameter and 7 ft (2130 mm) tube length.
• Type AFM size 33-96 (840-2440): This exchanger has a removable channel and cover
front head (A), two-pass shell (F) and fixed tube sheet bonnet-type rear head (M)
with 331/8
in (840 mm) inside diameter and 8ft (2440 mm) tube length.

Figure 3.1. TEMA-type designations for shell-and-tube heat exchangers. (Standards of Tubu-
lar Exchanger Manufacturers Association, 6th ed., 1978.)
In the above illustration the term single pass and two pass shell have been used. This
mean that the shell side fluid travels only one through the shell (single pass) or twice (two
pass shell). Two pass shell mean that the fluid enters at one end, travel to other end and
back to the end where it entered (making U-turn). Similarly there are multiple pases. To
be remembered is that the number of tube passes is equal to or greater than the number
of shell passes. Generally the multi shell and tube passes are usually designated by two
numerals separated by a hyphen, with the first numeral indication the number of shell
pass and the other stands for the tube passes. For example a one-shell pass and two tube
pass AEL exchanger will be written as 1-2 AEL. To be remembered is that this not an
TEMA standards. TEMA requires the number of shell and tube passes to be spelled out

3.1 TEMA Designations 31
as in the pervious examples. In a heat exchanger speciﬁcation sheet there is a space for
indicating the number of shell and tube passes. Another identiﬁcation of the shell and
tube heat exchanger is the number of shell passes. 1 shell pass, 2 shell pass, etc. This is
not a TEMA standardization. The tube passes can be equal to or greater than the shell
pass.

Table3.2.FeaturesofTEMAShell-and-Tube-TypeExchangers.
PackedInternalOutside
TypeFixedlantern-ringfloatingheadpackedPull-through
ofdesigntubesheetU-tubefloatinghead(splitbackingring)floatingheadfloatinghead
T.E.M.A.rear
-headtypeLorMorNUWSPT
Relativecostincreases
fromA(least
expensive)through
E(mostexpensive)BACEDE
Provisionfor
differentialexpansionExpansionIndividualtubesFloatingheadFloatingheadFloatingheadFloatinghead
jointinfreetoexpand
shell
RemovablebundleNoYesYesYesYesYes
Replacementbundle
possibleNoYesYesYesYesYes
Individualtubes
replaceableYesOnlythoseinYesYesYesYes
outsiderow
Tubecleaningby
chemicalsinside
andoutsideYesYesYesYesYesYes
Interiortube
cleaningmechanicallyYesSpecialtoolsrequiredYesYesYesYes
Exteriortube
cleaningmechanically:
TriangularpitchNoNoNoNoNoNo
SquarepitchNoYesYesYesYesYes
Hydraulic-jet
cleaning:
TubeinteriorYesSpecialtoolsrequiredYesYesYesYes
TubeexteriorNoYesYesYesYesYes
Doubletube
sheetfeasibleYesYesNoNoYesNo
NumberoftubepassesNopracticalAnyevenLimitedtooneNopracticalNopracticalNopractical
limitationsnumberpossibleortwopasseslimitationslimitationslimitations
Internalgaskets
eliminatedYesYesYesNoYesNo

3.2 Classification by construction STHE 33
3.2 Classification by construction STHE
Fig. 3.2 to Fig. 3.8 show details of the construction of the TEMA types of shell-and-tube
heat exchangers. These types are:
• Fixed tube sheet
• U-tube
• Floating head
3.2.1 Fixed tube sheet
Fixed-tube-sheet exchangers (Fig. 3.2) are used more often than any other type, and
the frequency of use has been increasing in recent years. The tube sheets are welded
to the shell. Usually these extend beyond the shell and serve as flanges to which the
tube-side headers are bolted. This construction requires that the shell and tube-sheet
materials be weldable to each other. When such welding is not possible, a blind-gasket
type of construction is utilized. The blind gasket is not accessible for maintenance or
replacement once the unit has been constructed. This construction is used for steam
surface condensers, which operate under vacuum.
Figure 3.2. Heat-exchanger-component nomenclature. Fixed tube heat sheet shell and tube
heat exchanger. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)
The tube-side header (or channel) may be welded to the tube sheet, as shown in Fig. 3.1
for type C and N heads. This type of construction is less costly than types B and M or
A and L and still offers the advantage that tubes may be examined and replaced without
disturbing the tube-side piping connections. There is no limitation on the number of
tube-side passes. Shell-side passes can be one or more, although shells with more than
two shell side passes are rarely used. Tubes can completely fill the heat-exchanger shell.
Clearance between the outermost tubes and the shell is only the minimum necessary
for fabrication. Between the inside of the shell and the baffles some clearance must be
provided so that baffles can slide into the shell. Fabrication tolerances then require some
additional clearance between the outside of the baffles and the outermost tubes. The edge
distance between the outer tube limit (OTL) and the baffle diameter must be sufficient
to prevent vibration of the tubes from breaking through the baffle holes. The outermost
tube must be contained within the OTL.
Clearances between the inside shell diameter and OTL are 13 mm (1/2 in) for 635-mm-
(25-in-) inside-diameter shells and up, 11 mm for 254- through 610-mm (10- through
24-in) pipe shells, and slightly less for smaller-diameter pipe shells.

Tubes can be replaced. Tube-side headers, channel covers, gaskets, etc., are accessible for
maintenance and replacement. Neither the shell-side baffle structure nor the blind gasket
is accessible. During tube removal, a tube may break within the shell. When this occurs,
it is most difficult to remove or to replace the tube. The usual procedure is to plug the
appropriate holes in the tube sheets.
Differential expansion between the shell and the tubes can develop because of differences
in length caused by thermal expansion. Various types of expansion joints are used to
eliminate excessive stresses caused by expansion. The need for an expansion joint is a
function of both the amount of differential expansion and the cycling conditions to be
expected during operation. A number of types of expansion joints are available (Fig. 3.3)
Figure 3.3. Expansion joints.
.
a Flat plates. Two concentric flat plates with a bar at the outer edges. The flat plates
can flex to make some allowance for differential expansion. This design is generally
used for vacuum service and gauge pressures below 103 kPa (15 lbf/in2). All welds
are subject to severe stress during differential expansion.
b Flanged-only heads. The flat plates are flanged (or curved). The diameter of these
heads is generally 203 mm (8 in) or more greater than the shell diameter. The
welded joint at the shell is subject to the stress referred to before, but the joint

connecting the heads is subjected to less stress during expansion because of the
curved shape.
c Flared shell or pipe segments. The shell may be flared to connect with a pipe
section, or a pipe may be halved and quartered to produce a ring.
d Formed heads. A pair of dished-only or elliptical or flanged and dished heads can
be used. These are welded together or connected by a ring. This type of joint is
similar to the flanged-only-head type but apparently is subject to less stress.
e Flanged and flued heads. A pair of flanged-only heads is provided with concentric
reverse flue holes. These heads are relatively expensive because of the cost of the
fluing operation. The curved shape of the heads reduces the amount of stress at the
welds to the shell and also connecting the heads.
f Toroidal. The toroidal joint has a mathematically predictable smooth stress pat-
tern of low magnitude, with maximum stresses at sidewalls of the corrugation and
minimum stresses at top and bottom. The foregoing designs were discussed as ring
expansion joints by Kopp and Sayre, Expansion Joints for Heat Exchangers (ASME
Misc. Pap., vol. 6, no. 211). All are statically indeterminate but are subjected
to analysis by introducing various simplifying assumptions. Some joints in current
industrial use are of lighter wall construction than is indicated by the method of
this paper.
g Bellows. Thin-wall bellows joints are produced by various manufacturers. These are
designed for differential expansion and are tested for axial and transverse movement
as well as for cyclical life. Bellows may be of stainless steel, nickel alloys, or copper.
(Aluminum, Monel, phosphor bronze, and titanium bellows have been manufac-
tured.) Welding nipples of the same composition as the heat-exchanger shell are
generally furnished. The bellows may be hydraulically formed from a single piece
of metal or may consist of welded pieces. External insulation covers of carbon steel
are often provided to protect the light-gauge bellows from damage. The cover also
prevents insulation from interfering with movement of the bellows (see h).
h Toroidal bellows. For high-pressure service the bellows type of joint has been modi-
fied so that movement is taken up by thin-wall small-diameter bellows of a toroidal
shape. Thickness of parts under high pressure is reduced considerably (see f ).
Improper handling during manufacture, transit, installation, or maintenance of the heat
exchanger equipped with the thin-wallbellows type or toroidal type of expansion joint can
damage the joint. In larger units these light-wall joints are particularly susceptible to
damage, and some designers prefer the use of the heavier walls of formed heads.
Chemical-plant exchangers requiring expansion joints most commonly have used the
flanged-and-flued-head type. There is a trend toward more common use of the light-
wall-bellows type.
3.2.2 U-Tube Heat Exchanger
Fig. 3.4 shows U-tube heat exchanger Type CFU. The tube bundle consists of a stationary
tube sheet, U tubes (or hairpin tubes), baffles or support plates, and appropriate tie rods
and spacers. The tube bundle can be removed from the heat-exchanger shell. A tube-side
header (stationary head) and a shell with integral shell cover, which is welded to the
shell, are provided. Each tube is free to expand or contract without any limitation being
placed upon it by the other tubes. The U-tube bundle has the advantage of providing
minimum clearance between the outer tube limit and the inside of the shell for any of
the removable-tube-bundle constructions. Clearances are of the same magnitude as for
fixed-tube-sheet heat exchangers. The number of tube holes in a given shell is less than

that for a fixed-tube-sheet exchanger because of limitations on bending tubes of a very
short radius.
Figure 3.4. Heat-exchanger-component nomenclature. U-tube heat exchanger. Type CFU.
(Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)
The U-tube design offers the advantage of reducing the number of joints. In high-pressure
construction this feature becomes of considerable importance in reducing both initial and
maintenance costs. The use of U-tube construction has increased significantly with the
development of hydraulic tube cleaners, which can remove fouling residues from both the
straight and the U-bend portions of the tubes. Rods and conventional mechanical tube
cleaners cannot pass from one end of the U tube to the other. Power-driven tube cleaners,
which can clean both the straight legs of the tubes and the bends, are available. Hydraulic
jetting with water forced through spray nozzles at high pressure for cleaning tube interiors
and exteriors of removal bundles is reported in the recent ASME publications.
U-tube can be used for high pressure and high temperature application like kettle reboiler,
evaporator, tank section heaters ,etc.
The tank suction heater, as illustrated in Fig. 3.5, contains a U-tube bundle. This design
is often used with outdoor storage tanks for heavy fuel oils, tar, molasses, and similar
fluids whose viscosity must be lowered to permit easy pumping. Uusally the tube-side
heating medium is steam. One end of the heater shell is open, and the liquid being heated
passes across the outside of the tubes. Pumping costs can be reduced without heating the
entire contents of the tank. Bare tube and integral low-fin tubes are provided with baffles.
Longitudinal fin-tube heaters are not baffled. Fins are most often used to minimize the
fouling potential in these fluids.

Figure 3.5. Heat-exchanger-component nomenclature. U-tube heat exchanger. Type CFU.
(Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)
Kettle-type reboilers, evaporators, etc. , are often U-tube exchangers with enlarged shell
sections for vapor-liquid separation (Fig.3.6). The U-tube bundle replaces the floating-
heat bundle of Fig. 3.4.
Figure 3.6. Kettle reboiler
The U-tube exchanger with copper tubes, cast-iron header, and other parts of carbon
steel is used for water and steam services in office buildings, schools, hospitals, hotels, etc.
Nonferrous tube sheets and admiralty or 90-10 copper-nickel tubes are the most frequently
used substitute materials. These standard exchangers are available from a number of
manufacturers at costs far below those of custombuilt process-industry equipment.
3.2.3 Floating Head Designs
In an effort to reduce thermal stresses and provide a means to remove the tube bundle
for cleaning, several floating rear head designs have been established. The simplest is a

Internal floating head (pull- through design) Fig3.9 design which allows the tube bundle to
be pulled entirely through the shell for service or replacement. In order to accommodate
the rear head bolt circle, tubes must be removed resulting in a less efficient use of shell
size. In addition, the missing tubes result in larger annular spaces and can contribute to
reduced flow across the effective tube surface, resulting in reduced thermal performance.
Some designs include sealing strips installed in the shell to help block the bypass steam.
Another floating head design that partially addresses the above disadvantages is a split-
ring floating head. Here the floating head bonnet is bolted to a split backing ring instead
of the tube sheet. This eliminates the bolt circle diameter and allows a full complement
of tubes to fill the shell. This construction is more expensive than a common pull through
design, but is in wide use in petrochemical applications. For applications with high
pressures or temperatures, or where more positive sealing between the fluids is desired,
the pull-through design should be specified.
Two other types, the outside packed lantern ring and the outside packed stuffing box
designs offer less positive sealing against leakage to the atmosphere than the pull though
or split ring designs, but can be configured for single tube pass duty. More details about
the various types of floating head shell and tube heat exchanger is given the following
sections
Packed-Lantern-Ring Exchanger: (Fig. 3.7 ) This construction is the least costly
of the straight-tube removable bundle types. The shell- and tube-side fluids are each
contained by separate rings of packing separated by a lantern ring and are installed at the
floating tube sheet. The lantern ring is provided with weep holes. Any leakage passing
the packing goes through the weep holes and then drops to the ground. Leakage at the
packing will not result in mixing within the exchanger of the two fluids. The width of the
floating tube sheet must be great enough to allow for the packings, the lantern ring, and
differential expansion. Sometimes a small skirt is attached to a thin tube sheet to provide
the required bearing surface for packings and lantern ring. The clearance between the
outer tube limit and the inside of the shell is slightly larger than that for fixed-tube-sheet
and U-tube exchangers.
The use of a floating-tube-sheet skirt increases this clearance. Without the skirt the
clearance must make allowance for tubehole distortion during tube rolling near the outside
edge of the tube sheet or for tube-end welding at the floating tube sheet.
The packed-lantern-ring construction is generally limited to design temperatures below
191◦
C (375◦
F) and to the mild services of water, steam, air, lubricating oil, etc. Design
gauge pressure does not exceed 2068 kPa (300 lbf/in2
) for pipe shell exchangers and is
limited to 1034 kPa (150 lbf/in2
) for 610- to 1067-mm- (24- to 42-in-) diameter shells.

Figure 3.7. Heat-exchanger-component nomenclature. Exchanger with packed floating tube
sheet and lantern ring. Type AJW. External floating head design. (Standard of Tubular Ex-
changer Manufacturers Association, 6th ed., 1978.)
Outside-Packed Floating-Head Exchanger: (Fig. 3.8) The shell-side fluid is con-
tained by rings of packing, which are compressed within a stuffing box by a packing
follower ring. This construction was frequently used in the chemical industry, but in
recent years usage has decreased. The removable-bundle construction accommodates dif-
ferential expansion between shell and tubes and is used for shell-side service up to 4137
kPa gauge pressure (600 lbf/in2) at 316◦
C (600◦
F).
Figure 3.8. Heat-exchanger-component nomenclature. Outside-packed floating-head ex-
changer. Type AEP. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)
There are no limitations upon the number of tube-side passes or upon the tube-side
design pressure and temperature. The outside-packed floating-head exchanger was the
most commonly used type of removable- bundle construction in chemical-plant service.
The floating-tube-sheet skirt, where in contact with the rings of packing, has fine machine
finish. A split shear ring is inserted into a groove in the floating-tube-sheet skirt. A slip-
on backing flange, which in service is held in place by the shear ring, bolts to the external
floating- head cover. The floating-head cover is usually a circular disk. With an odd
number of tube-side passes, an axial nozzle can be installed in such a floating- head cover.
If a side nozzle is required, the circular disk is replaced by either a dished head or a channel
barrel (similar to Fig. 11-36f ) bolted between floating-head cover and floating-tube-sheet
skirt. The outer tube limit approaches the inside of the skirt but is farther removed from
the inside of the shell than for any of the previously discussed constructions. Clearances

between shell diameter and bundle OTL are 22 mm (7.8 in) for small-diameter pipe shells,
44 mm (1e in) for large-diameter pipe shells, and 58 mm (2g in) for moderatediameter
plate shells.
Internal Floating-Head Exchanger: (Fig. 3.9) The internal floating-head design
is used extensively in petroleum-refinery service, but in recent years there has been a
decline in usage. The tube bundle is removable, and the floating tube sheet moves (or
floats) to accommodate differential expansion between shell and tubes. The outer tube
limit approaches the inside diameter of the gasket at the floating tube sheet. Clearances
(between shell and OTL) are 29 mm for pipe shells and 37 mm for moderatediameter plate
shells. A split backing ring and bolting usually hold the floating-head cover at the floating
tube sheet. These are located beyond the end of the shell and within the larger-diameter
shell cover. Shell cover, split backing ring, and floating-head cover must be removed before
the tube bundle can pass through the exchanger shell. With an even number of tube-side
passes the floating-head cover serves as return cover for the tube-side fluid. With an odd
number of passes a nozzle pipe must extend from the floating-head cover through the shell
cover. Provision for both differential expansion and tube-bundle removal must be made.
Figure 3.9. Heat-exchanger-component nomenclature. Internal floating head (pull- through
design). Type AES. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)
Figure 3.10. Heat-exchanger-component nomenclature. Exchanger with packed floating tube
sheet and lantern ring. Type AES. (Standard of Tubular Exchanger Manufacturers Association,
6th ed., 1978.)
Pull-Through Floating-Head Exchanger: (Fig. 3.12) Construction is similar to that
of the internal-floating-head split-backing ring exchanger except that the floating-head

3.3 Shell Constructions 41
cover bolts directly to the floating tube sheet. The tube bundle can be withdrawn from
the shell without removing either shell cover or floating-head cover. This feature reduces
maintenance time during inspection and repair.
The large clearance between the tubes and the shell must provide for both the gasket
and the bolting at the floating-head cover. This clearance is about 2 to 2.5 times that
required by the split-ring design. Sealing strips or dummy tubes are often installed to
reduce bypassing of the tube bundle.
Figure 3.11. Heat-exchanger-component nomenclature. Kettle-type floating-head reboiler.
Type AKT. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)
3.3 Shell Constructions
• The most common TEMA shell type is the E shell as it is most suitable for most
industrial process cooling applications. However, for certain applications, other
shells offer distinct advantages. For example, the TEMA-F shell design provides
for a longitudinal flow plate to be installed inside the tube bundle assembly. This
plate causes the shell fluid to travel down one half of the tube bundle, then down
the other half, in effect producing a counter-current flow pattern which is best for
heat transfer. This type of construction can be specified where a close approach
temperature is required and when the flow rate permits the use of one half of the
shell at a time. In heat recovery applications, or where the application calls for
increased thermal length to achieve effective overall heat transfer, shells can be
installed with the flows in series. Up to six shorter shells in series is common and
results in counter-current flow close to performance as if one long shell in a single
pass design were used.
• TEMA G and H shell designs are most suitable for phase change applications where
the bypass around the longitudinal plate and counter-current flow is less impor-
tant than even flow distribution. In this type of shell, the longitudinal plate offers
better flow distribution in vapor streams and helps to flush out non-condensable.
They are frequently specified for use in horizontal thermosiphon reboilers and total
condensers.
• TEMA J Shells are typically specified for phase change duties where significantly
reduced shell side pressure drops are required. They are commonly used in stacked
sets with the single nozzles used as the inlet and outlet. A special type of J-shell
is used for flooded evaporation of shell side fluids. A separate vapor disengagement
vessel without tubes is installed above the main J shell with the vapor outlet at the
top of this vessel. The

• TEMA K shell, also termed a kettle reboiler, is specified when the shell side stream
will undergo vaporization. The liquid level of a K shell design should just cover the
tube bundle, which fills the smaller diameter end of the shell. This liquid level is
controlled by the liquid flowing over a weir at the far end of the entrance nozzle. The
expanded shell area serves to facilitate vapor disengagement for boiling liquid in the
bottom of the shell. To insure against excessive liquid carry-though with the vapor
stream, a separate vessel as described above is specified. Liquid carry-through can
also be minimized by installing a mesh demister at the vapor exit nozzle. U-bundles
are typically used with K shell designs. K shells are expensive for high pressure
vaporization due to shell diameter and the required wall thickness.
• The TEMA X shell, or crossflow shell is most commonly used in vapor condensing
applications, though it can also be used effectively in low pressure gas cooling or
heating. It produces a very low shell side pressure drop, and is therefore most
suitable for vacuum service condensing. In order to assure adequate distribution
of vapors, X-shell designs typically feature an area free of tubes along the top of
the exchanger. It is also typical to design X shell condensers with a flow area at
the bottom of the tube bundle to allow free condensate flow to the exit nozzle.
Careful attention to the effective removal of non-condensables is vital to X-shell
constructions.
3.4 Tube side construction
3.4.1 Tube-Side Header:
The tube-side header (or stationary head) contains one or more flow nozzles.
• The bonnet (Fig. 3.1B) bolts to the shell. It is necessary to remove the bonnet in
order to examine the tube ends. The fixed-tubesheet exchanger of Fig. 3.1b has
bonnets at both ends of the shell.
• The channel (Fig. 3.1A) has a removable channel cover. The tube ends can be
examined by removing this cover without disturbing the piping connections to the
channel nozzles. The channel can bolt to the shell as shown in Fig. 3.1a and c.
The Type C and Type N channels of Fig. 3.1 are welded to the tube sheet. This
design is comparable in cost with the bonnet but has the advantages of permitting
access to the tubes without disturbing the piping connections and of eliminating a
gasketed joint.
• Special High-Pressure Closures (Fig. 3.1D) The channel barrel and the tube sheet
are generally forged. The removable channel cover is seated in place by hydrostatic
pressure, while a shear ring subjected to shearing stress absorbs the end force. For
pressures above 6205 kPa (900 lbf/in2) these designs are generally more economical
than bolted constructions, which require larger flanges and bolting as pressure in-
creases in order to contain the end force with bolts in tension. Relatively light-gauge
internal pass partitions are provided to direct the flow of tube-side fluids but are
designed only for the differential pressure across the tube bundle.
3.4.2 Tube-Side Passes
Most exchangers have an even number of tube-side passes. The fixed-tube-sheet exchanger
(which has no shell cover) usually has a return cover without any flow nozzles as shown in
Fig. 3.1M; Types L and N are also used. All removable-bundle designs (except for the U
tube) have a floating-head cover directing the flow of tube-side fluid at the floating tube
sheet.

3.4 Tube side construction 43
3.4.3 Tubes Type
There are different type of tubes used in heat exchangers. These are
1. Plain tube
(a) Straight tube
(b) U-tube with a U-bend
(c) Coiled tubes
2. Finned tube
3. Duplex or bimetallic tube. These tube are in reality two tube of different materials,
one closely fitted over the other with no gap between them. They are made by
drawing the outer tube onto the inner one or by shrink fitting. These are used
where corrosive nature of the tube side fluid is such that no one metal or alloy is
compatible with fluids.
4. Enhanced surface tube
1. Plain tube
Standard heat-exchanger tubing is (1/4, 3/8, 1/2, 5/8, 3/4, 1, 1 1/4, 1 1/2 inch in
outside diameter (1 inch= 25.4 mm). Wall thickness is measured in Birmingham
wire gauge (BWG) units. The most commonly used tubes in chemical plants and
petroleum refineries are 19- and 25-mm (3/4- and 1-in) outside diameter. Standard
tube lengths are 8, 10, 12, 16, and 20 ft, with 20 ft now the most common ( 1 ft=
0.3048 m).
Manufacturing tolerances for steel, stainless-steel, and nickel alloy tubes are such
that the tubing is produced to either average or minimum wall thickness. Seamless
carbon steel tube of minimum wall thickness may vary from 0 to 20 percent above the
nominal wall thickness. Average-wall seamless tubing has an allowable variation of
plus or minus 10 percent. Welded carbon steel tube is produced to closer tolerances
(0 to plus 18 percent on minimum wall; plus or minus 9 percent on average wall).
Tubing of aluminum, copper, and their alloys can be drawn easily and usually is
made to minimum wall specifications.
Common practice is to specify exchanger surface in terms of total external square
feet of tubing. The effective outside heat-transfer surface is based on the length of
tubes measured between the inner faces of tube sheets. In most heat exchangers
there is little difference between the total and the effective surface. Significant
differences are usually found in high-pressure and double-tube-sheet designs.
Tube thickness The tube should be able to stand:
(a) pressure on the inside and out side of the tube
(b) temperature on both the sides
(c) thermal stress due to the differential expansion of the shell and the tube bundle
(d) corrosive nature of both the shell-side and the tube side fluid
The tube thickness is given a function of the tube out side diameter in accordance
with B.W.G.

Figure 3.12. Tube thickness
2. Finned tube: As the name implies, finned tube have fins to the tubular surface.
Fins can be longtiudinal, radial or helical and may be on the outside or inside or on
both sides of the tube. Fig. 5.7shows some of the commonly used fins. The fins are
generally used when at least one of the fluid is gas.

3.4 Tube side construction 45
Figure 3.13. Examples of extended surfaces on one or both sides. (a) Radial fins. (b) Serrated
radial fins. (c) Studded surface. (d) Joint between tubesheet and low fin tube with three times
bare surface. (e) External axial fins. ( f ) Internal axial fins. (9) Finned surface with internal
spiral to promote turbulence. (h) Plate fins on both sides. (i) Tubes and plate fins.
(a) Integrally finned tube, which is available in a variety of alloys and sizes, is
being used in shell-and-tube heat exchangers. The fins are radially extruded
from thick-walled tube to a height of 1.6 mm (1/16 in) spaced at 1.33 mm (19
fins per inch) or to a height of 3.2 mm (1/8 in) spaced at 2.3 mm (11 fins per
inch). External surface is approximately 2 1/2 times the outside surface of a
bare tube with the same outside diameter. Also available are 0.93-mm- (0.037-
in-) high fins spaced 0.91 mm (28 fins per inch) with an external surface about
3.5 times the surface of the bare tube. Bare ends of nominal tube diameter are
provided, while the fin height is slightly less than this diameter. The tube can
be inserted into a conventional tube bundle and rolled or welded to the tube
sheet by the same means, used for bare tubes. An integrally finned tube rolled

into a tube sheet with double serrations and flared at the inlet is shown in
Fig. 11-39. Internally finned tubes have been manufactured but have limited
application.
(b) Longitudinal fins are commonly used in double-pipe exchangers upon the
outside of the inner tube. U-tube and conventional removable tube bundles
are also made from such tubing. The ratio of external to internal surface
generally is about 10 or 15:1.
(c) Transverse fins upon tubes are used in low-pressure gas services. The primary
application is in air-cooled heat exchangers (as discussed under that heading),
but shell-and-tube exchangers with these tubes are in service.
3. Bimetallic Tubes When corrosive requirements or temperature conditions do not
permit the use of a single alloy for the tubes, bimetallic (or duplex) tubes may be
used. These can be made from almost any possible combination of metals. Tube
sizes and gauges can be varied. For thin gauges the wall thickness is generally
divided equally between the two components. In heavier gauges the more expensive
component may comprise from a fifth to a third of the total thickness.
The component materials comply with applicable ASTM specifications, but after
manufacture the outer component may increase in hardness beyond specification
limits, and special care is required during the tube-rolling operation. When the
harder material is on the outside, precautions must be exercised to expand the
tube properly. When the inner material is considerably softer, rolling may not be
practical unless ferrules of the soft material are used.
In order to eliminate galvanic action the outer tube material may be stripped from
the tube ends and replaced with ferrules of the inner tube material. When the end
of a tube with a ferrule is expanded or welded to a tube sheet, the tube-side fluid
can contact only the inner tube material, while the outer material is exposed to the
shell-side fluid. Bimetallic tubes are available from a small number of tube mills
and are manufactured only on special order and in large quantities.
4. Enhance surface These kind of tubes enhance the heat transfer coefficient (Fig.
5.7h,i). This may be achieved by two techniques.
(a) The surface is contoured or grooved in a variety of ways forming valley and
ridges. These are applicable in condenser and.
(b) The surface is prepared with special coating to provide a large number of
nucleation sites for use in boiling operations.
3.4.4 Tube arrangement
The tubes in an exchanger are usually arranged in an equilateral triangular, aquare or
rotated square pattern see fig.3.14.
The triangular and rotated square pattern give higher heat transfer rates, but at the
expenses of higher pressure drop than the the square pattern. Square or rotated square
are used for hihger fouling fluid, where it is necessary to mechanically clean the outside
of the tubes. The recommend tube pitch is Pt = 1.25do. Where square pattern is used
for easer of cleaning, the recommended minimum clearance between the tubes is 0.25 in
(6.4 mm)

3.5 Shell side construction 47
pt
do
Square pitch
pt
Equilateral triangular pitch
pt
d
o
Rotaed square
Flow
Figure 3.14. Tube patterns.
3.4.5 Tube side passes
The fluid in the tube is usually directed to flow back and forth in a number of passes
through groups of tube arranged in parallel to increase the length of the flow path. The
number of passes is selected to give the required side design velocity. Exchangers are built
form one to up to 16 passes. The tube are arranged into the number of passes required by
dividing up the exchanger headers (channels) with partition plates (pass partition) The
arrangement of the pass partition for 2,4 and 6 are shown in fig.3.19
1
2
1
2
3
4
1
2 3
45
6
Two tube passes
Four tube passes
Six tube passes
1
2 3
45
6
Figure 3.15. Tube arrangement: showing pass-partitions in headers.
3.5 Shell side construction
3.5.1 Shell Sizes
Heat-exchanger shells are generally made from standard- wall steel pipe in sizes up to
305-mm (12-in) diameter; from 9.5-mm (3/8 in) wall pipe in sizes from 356 to 610 mm
(14 to 24 in); and from steel plate rolled at discrete intervals in larger sizes. Clearances
between the outer tube limit and the shell are discussed elsewhere in connection with the
different types of construction.

3.5.2 Shell-Side Arrangements
1. The one-pass shell (Fig. 3.1E) is the most commonly used arrangement. Con-
densers from single component vapors often have the nozzles moved to the center
of the shell for vacuum and steam services. Solid longitudinal baffle is provided to
form a two-pass shell (Fig. 3.1F). It may be insulated to improve thermal efficiency.
(See further discussion on baffles).
2. A two-pass shell can improve thermal effectiveness at a cost lower than for two
shells in series.
3. For split flow (Fig. 3.1G), the longitudinal baffle may be solid or perforated. The
latter feature is used with condensing vapors.
4. double-split-flow design is shown in Fig. 3.1H. The longitudinal baffles may be
solid or perforated.
5. The divided flow design (Fig. 3.1J), mechanically is like the one-pass shell ex-
cept for the addition of a nozzle. Divided flow is used to meet low-pressure-drop
requirements. The kettle reboiler is shown in Fig. 3.1K. When nucleate boiling is
to be done on the shell-side, this common design provides adequate dome space for
separation of vapor and liquid above the tube bundle and surge capacity beyond
the weir near the shell cover.
3.6 Baffles and tube bundles
3.6.1 The tube bundle
Tube bundle is the most important part of a tubular heat exchanger. The tubes generally
constitute the most expensive component of the exchanger and are the one most likely to
corrode. Tube sheets, baffles, or support plates, tie rods, and usually spacers complete
the bundle.
3.6.2 Baffle
Baffles are used to direct the side and tube side flows so that the fluid velocity is increased
to obtain higher heat transfer rate and reduce fouling deposits. In horizontal units baffle
are used to provide support against sagging and vibration damage. There are different
types of baffles:
1. segemntal
2. disc and doughnut
3. orifice
4. rod type
5. nest type
6. longitudinal
7. impingment
1. Segmental Baffles Segmental or cross-flow baffles are standard. Single, double,
and triple segmental baffles are used. Baffle cuts are illustrated in Fig. 3.16a. The
double segmental baffle reduces crossflow velocity for a given baffle spacing. The
triple segmental baffle reduces both cross-flow and long-flow velocities and has been
identified as the window-cut baffle.

3.6 Baffles and tube bundles 49
a
b
c
d
Figure 3.16. Types of baffle used in shell and tube heat exchanger. (a) Segmental. (b)
Segmental and strip. (c) Disc and doughnut. (d) Oriffice.
Minimum baffle spacing is generally one-fifth of the shell diameter and not less
than 50.8 mm (2 in). Maximum baffle spacing is limited by the requirement to
provide adequate support for the tubes. The maximum unsupported tube span
in inches equals 74d0.75
(where d is the outside tube diameter in inches). The
unsupported tube span is reduced by about 12 percent for aluminum, copper, and
their alloys.
Baffles are provided for heat-transfer purposes. When shell-side baffles are not
required for heat-transfer purposes, as may be the case in condensers or reboilers,
tube supports are installed.
Maximum baffle cut is limited to about 45 percent for single segmental baffles so
that every pair of baffles will support each tube. Tube bundles are generally pro-
vided with baffles cut so that at least one row of tubes passes through all the baffles
or support plates. These tubes hold the entire bundle together. In pipe-shell ex-
changers with a horizontal baffle cut and a horizontal pass rib for directing tube

side flow in the channel, the maximum baffle cut, which permits a minimum of one
row of tubes to pass through all baffles, is approximately 33 percent in small shells
and 40 percent in larger pipe shells.
Maximum shell-side heat-transfer rates in forced convection are apparently obtained
by cross-flow of the fluid at right angles to the tubes. In order to maximize this
type of flow some heat exchangers are built with segmental-cut baffles and with no
tubes in the window (or the baffle cutout). Maximum baffle spacing may thus equal
maximum unsupported-tube span, while conventional baffle spacing is limited to
one-half of this span.
The maximum baffle spacing for no tubes in the window of single segmental baffles
is unlimited when intermediate supports are provided. These are cut on both sides
of the baffle and therefore do not affect the flow of the shell-side fluid. Each support
engages all the tubes; the supports are spaced to provide adequate support for the
tubes.
2. Rod Baffles Rod or bar baffles (fig. 3.17) have either rods or bars extending
through the lanes between rows of tubes. A baffle set can consist of a baffle with
rods in all the vertical lanes and another baffle with rods in all the horizontal lanes
between the tubes. The shell-side flow is uniform and parallel to the tubes. Stagnant
areas do not exist.
One device uses four baffles in a baffle set. Only half of either the vertical or the
horizontal tube lanes in a baffle have rods. The new design apparently provides a
maximum shell-side heat-transfer coefficient for a given pressure drop.
Figure 3.17. Rod baffles.
3. Impingement Baffle The tube bundle is customarily protected against impinge-
ment by the incoming fluid at the shell inlet nozzle when the shell-side fluid is at a
high velocity, is condensing, or is a twophase fluid. Minimum entrance area about
the nozzle is generally equal to the inlet nozzle area. Exit nozzles also require ade-
quate area between the tubes and the nozzles. A full bundle without any provision
for shell inlet nozzle area can increase the velocity of the inlet fluid by as much as
300 percent with a consequent loss in pressure.
Impingement baffles are generally made of rectangular plate, although circular plates
(Fig. 3.18) are more desirable. Rods and other devices are sometimes used to
protect the tubes from impingement. In order to maintain a maximum tube count

the impingement plate is often placed in a conical nozzle opening or in a dome cap
above the shell.
Impingement baffles or flow-distribution devices are recommended for axial tube-
side nozzles when entrance velocity is high.
(a)
(B)
(c) (d)
Figure 3.18. Impingment baffless;(a)Flat plate (b)curved plate (c)expanded or flared nozzle
(d) jacket type.
4. Longitudinal Flow Baffles In fixed-tube-sheet construction with multipass shells,
the baffle is usually welded to the shell and positive assurance against bypassing
results. Removable tube bundles have a sealing device between the shell and the
longitudinal baffle. Flexible light-gauge sealing strips and various packing devices
have been used. Removable U-tube bundles with four tube-side passes and two
shell-side passes can be installed in shells with the longitudinal baffle welded in
place.
In split-flow shells the longitudinal baffle may be installed without a positive seal
at the edges if design conditions are not seriously affected by a limited amount of
bypassing.
Fouling in petroleum-refinery service has necessitated rough treatment of tube bun-
dles during cleaning operations. Many refineries avoid the use of longitudinal baffles,
since the sealing devices are subject to damage during cleaning and maintenance
operations.
3.6.3 Vapor Distribution
Relatively large shell inlet nozzles, which may be used in condensers under low pressure
or vacuum, require provision for uniform vapor distribution.
3.6.4 Tube-Bundle Bypassing
Shell-side heat-transfer rates are maximized when bypassing of the tube bundle is at a
minimum. The most significant bypass stream is generally between the outer tube limit
and the inside of the shell. The clearance between tubes and shell is at a minimum for
fixed-tube-sheet construction and is greatest for straight-tube removable bundles. Ar-
rangements to reduce tube-bundle bypassing include:

1. Dummy tubes. These tubes do not pass through the tube sheets and can be
located close to the inside of the shell.
2. Tie rods with spacers. These hold the baffles in place but can be located to
prevent bypassing.
3. Sealing strips. These longitudinal strips either extend from baffle to baffle or may
be inserted in slots cut into the baffles.
4. Dummy tubes or tie rods with spacers may be located within the pass partition
lanes (and between the baffle cuts) in order to ensure maximum bundle penetration
by the shell-side fluid.
When tubes are omitted from the tube layout to provide entrance area about an
impingement plate, the need for sealing strips or other devices to cause proper
bundle penetration by the shell-side fluid is increased.
3.6.5 Tie Rods and Spacers
Tie rods are used to hold the baffles in place with spacers, which are pieces of tubing or
pipe placed on the rods to locate the baffles. Occasionally baffles are welded to the tie
rods, and spacers are eliminated. Properly located tie rods and spacers serve both to hold
the bundle together and to reduce bypassing of the tubes.
In very large fixed-tube-sheet units, in which concentricity of shells decreases, baffles are
occasionally welded to the shell to eliminate bypassing between the baffle and the shell.
Metal baffles are standard. Occasionally plastic baffles are used either to reduce corrosion
or in vibratory service, in which metal baffles may cut the tubes.
Tube plate
baffle
Spacer
Rods
Figure 3.19. Baffle spacers and tie rods.
3.6.6 Tubesheets
Tubesheets are usually made from a round flat piece of metal with holes drilled for the
tube ends in a precise location and pattern relative to one another. Tube sheet materials
range as tube materials. Tubes are attached to the tube sheet by pneumatic or hydraulic
pressure or by roller expansion. Tube holes can be drilled and reamed and can be machined
with one or more grooves. This greatly increases the strength of the tube joint.

0.4mm
3 mm
a
b c
Figure 3.20. Tube sheet joint
The tubesheet is in contact with both fluids and so must have corrosion resistance al-
lowances and have metalurgical and electrochemical properties appropriate for the fluids
and velocities. Low carbon steel tube sheets can include a layer of a higher alloy metal
bonded to the surface to provide more effective corrosion resistance without the expense
of using the solid alloy. The tube hole pattern or pitch varies the distance from one tube
to the other and angle of the tubes relative to each other and to the direction of flow. This
allows the manipulation of fluid velocities and pressure drop, and provides the maximum
amount of turbulance and tube surface contact for effective heat transfer. Where the
tube and tube sheet materials are joinable, weldable metals, the tube joint can be further
strengthened by applying a seal weld or strength weld to the joint. A strength weld has
a tube slightly reccessed inside the tube hole or slightly extended beyond the tube sheet.
The weld adds metal to the resulting lip. A seal weld is specified to help prevent the
shell and tube liquids from intermixing. In this treatment, the tube is flush with the tube
sheet surface. The weld does not add metal, but rather fuses the two materials. In cases
where it is critical to avoid fluid intermixing, a double tube sheet can be provided. In this
design, the outer tube sheet is outside the shell circuit, virtually eliminating the chance
of fluid intermixing. The inner tube sheet is vented to atmosphere so any fluid leak is
easily detected.
Mechanisms of attaching tubes to tube sheet
• Rolled Tube Joints Expanded tube-to-tube-sheet joints are standard. Properly
rolled joints have uniform tightness to minimize tube fractures, stress corrosion,
tube-sheet ligament pushover and enlargement, and dishing of the tube sheet. Tubes
are expanded into the tube sheet for a length of two tube diameters, or 50 mm (2
in), or tube-sheet thickness minus 3 mm (1/8 in). Generally tubes are rolled for the
last of these alternatives. The expanded portion should never extend beyond the
shell-side face of the tube sheet, since removing such a tube is extremely difficult.
Methods and tools for tube removal and tube rolling were discussed by John, 1959.
Tube ends may be projecting, flush, flared, or beaded (listed in order of usage). The
flare or bell-mouth tube end is usually restricted to water service in condensers and
serves to reduce erosion near the tube inlet.
For moderate general process requirements at gauge pressures less than 2058 kPa
(300 lbf/in2) and less than 177◦
C (350◦
F), tube-sheet holes without grooves are
standard. For all other services with expanded tubes at least two grooves in each
tube hole are common. The number of grooves is sometimes changed to one or three
in proportion to tube-sheet thickness.
• Expanding the tube into the grooved tube holes provides a stronger joint but
results in greater difficulties during tube removal (see Fig. 3.20a).

• Welded Tube Joints When suitable materials of construction are used, the tube
ends may be welded to the tube sheets. Welded joints may be seal-welded for addi-
tional tightness beyond that of tube rolling or may be strength-welded. Strength-
welded joints have been found satisfactory in very severe services. Welded joints
may or may not be rolled before or after welding (see Fig. 3.20b).
The variables in tube-end welding were discussed in two unpublished papers [39] and
[119]. Tube-end rolling before welding may leave lubricant from the tube expander in
the tube hole. Fouling during normal operation followed by maintenance operations
will leave various impurities in and near the tube ends. Satisfactory welds are rarely
possible under such conditions, since tube-end welding requires extreme cleanliness
in the area to be welded.
• Tube expansion after welding has been found useful for low and moderate pres-
sures. In high-pressure service tube rolling has not been able to prevent leakage
after weld failure.
• Double-Tube-Sheet Joints This design prevents the passage of either ﬂuid into
the other because of leakage at the tube-to-tubesheet joints, which are generally the
weakest points in heat exchangers. Any leakage at these joints admits the ﬂuid to
the gap between the tube sheets. Mechanical design, fabrication, and maintenance
of doubletube-sheet designs require special consideration (see Fig. 3.20c).

55
4 Basic Design Equations of Heat Exchangers
There are two types of design problems: sizing and rating. In sizing the main objective
is to find the geometry of the heat exchanger. Rating is to find the duty or performance
for a given geometry.
RATING SIZING
Given: Geometry Given: Q(duty)
mh, Ch, Th1, ∆ph mh, Ch, Th1, ∆ph
mc, Cc, Tc1, ∆pc mc, Cc, Tc1, ∆pc
Find: Q(Duty) Find: Geometry
The are three design approaches generally used in the design of heat exchanger. These
are
• LMTD-method,
• NTU-ε-method and
• θ-method.
These notation are explained in the respective sections.
4.1 LMTD-Method
Assumptions
• Steady state flow (mh, mc)
• Constant overall heat transfer coefficient (U)
• Constant specific heat (Cph, Cpc)
• negligible heat loss to surrounding
Heat Transfer (or rate equation)
Q = UA∆TlmF (4.1)
where
Q = heat transferred per unit time W (duty)
U = overall heat transfer coefficient
A = heat transfer area
∆Tlm = logarithmic mean temperature difference
F = temperature correction factor

56 4 Basic Design Equations of Heat Exchangers
4.1.1 Logarithmic mean temperature different
∆Tlm =
∆T2 − ∆T1
ln(∆T2/∆T1)
(4.2)
The temperature difference ∆T1, ∆T2 for different tube heat exchanger are defined below:
ThoThi
Tci Tco
Thi
Tho
Tco
Tci
∆T1 ∆T2
Cocurrent
ThoThi
Tco Tci
Thi
Tho
Tco
Tci
∆T1
∆T2
Counter current
Thi
Tho
Tco
Tci
∆T1
Thi
Tho
Tci
Tco
Tc
Shell and Tube
Figure 4.1. Temperature distribution
∆T1 ∆T2
Cocurrent Thi − Tci Tho − Tco
Counter current Thi − Tco Tho − Tci
Shell and tube Thi − Tco Tho − Tci
Plate heat exchanger Thi − Tco Tho − Tci
Example 1 water at a rate of 68 kg/min is heated from 35 to 65 o
C by an oil having a
specific heat of 1.9 kJ/kg o
C. The oil enters the exchanger at 110 o
C and leaves at 75 o
C.
Calculate the logarithmic mean temperature difference for
1. counter current
2. co-current
Solution

4.1 LMTD-Method 57
ThoThi
Tci Tco
Thi=110 Co
Tho=75 Co
Tco=65 Co
Tci= 35 Co
∆T =751 ∆T =10 C2
o
Cocurrent
ThoThi
Tco Tci
∆T1=45 Co
∆T C2=40o
Counter current
Thi=110 Co
Tci= 35 Co
Tci= 35 Co
Tco=65 Co
Tho=75 Co
Figure 4.2. Temperature distribution
1. counter current (see Fig.4.2)
∆Tlm =
∆T2 − ∆T1
ln(∆T2/∆T1)
=
10 − 75
ln(10/75)
= 32.26o
C (4.3)
2. co-current (see Fig.4.2)
∆Tlm =
∆T2 − ∆T1
ln(∆T2/∆T1)
=
40 − 45
ln(40/45)
= 42.45o
C (4.4)
4.1.2 Correction Factor
• For double pipe heat exchanger
F = 1 (4.5)
• Shell and tube heat exchanger. For a 1 shell 2 tube pass exchanger the correction
factor is given by:
F =
(R2 + 1) ln 1−S
1−RS
(R − 1) ln



2−S R+1−
√
(R2+1)
2−S R+1−
√
(R2+1)



(4.6)
where
R =
T1 − T2
t2 − t1
, S =
t2 − t1
T1 − t1
(4.7)
or in words
R =
Range of shell fluid
Range of tube fluid
, S =
Range of tube fluid
Maximum temperature difference
(4.8)
the derivation of the equation 4.6 is given by Kern (1950). The equation can be
used for any exchanger with an even number of tube passes and is plotted in Fig.4.4.
The correction factor for 2 shell passes and 4 or multiple of 4 tube passes is
F =
R2+1
2(R−1)
ln 1−S
1−RS
ln
2/S−1−R+(2/S)
√
(1−S)(1−RS)+
√
R2+1
2/S−1−R+(2/S)
√
(1−S)(1−RS)−
√
R2+1
(4.9)
These equations are plotted on ﬁg.4.4

Example 1 For example calculate the correction factor for
1. 1-2 shell and tube heat exchanger and
2. 2-4 shell and tube heat exchanger
using the equation and the graph.
T1 = 35o
C, T2 = 65o
C, t1 = 110o
C, t2 = 75o
C
R =
T1 − T2
t2 − t1
=
35 − 65
75 − 110
= 0.86, S =
t2 − t1
T1 − t1
=
75 − 110
35 − 110
= 0.467 (4.10)
From the graph of ﬁg.4.4
1. for 1-2 shell and tube heat exchanger F=0.92
2. for 2-4 shell and tube heat exchanger F=0.98
T2
t1
t2
1-2 Shell and Tube
T1 t1
T1
T2
t2
2-4 Shell and Tube
Figure 4.3. Temperature distribution for 1-2 and 2-4 shell and tube heat exchanger

4.1 LMTD-Method 59
Figure 4.4. Temperature correction factor: one shell, 2 shell pass, divide flow shell and split
flow shell and cross flow
4.1.3 Overall heat transfer coefficient
Typical values of the overall heat transfer coefficient for various types of heat exchnager
are given in . More expensive data can be found in in
The determination of U is often tedious and needs data not yet available in preliminary
stages of the design. Therefore, typical values of U are useful for quickly estimating the
required surface area. The literature has many tabulations of such typical coefficients for
commercial heat transfer services.
Following is a table 4.1 with values for different applications and heat exchanger types.
More values can be found in the books as [29],[127], [113], [79], [93] and [14]
The ranges given in the table are an indication for the order of magnitude. Lower values
are for unfavorable conditions such as lower flow velocities, higher viscosities, and addi-
tional fouling resistances. Higher values are for more favorable conditions. Coefficients
of actual equipment may be smaller or larger than the values listed. Note that the val-
ues should not be used as a replacement of rigorous methods for the final design of heat
exchangers, although they may serve as a useful check on the results obtained by these

methods.
Table 4.1. Typical overall coefficient
Hot Fluid Cold fluid U (W/m2 o
C)
Heat exchangers
Water Water 800-1500
Organic solvents organic solvent 100-300
light oils light oils 100-400
heavy oils heavy oils 50-300
Gases gass 10-50
Coolers
Organic solvents water 250-750
light oils water 350-900
heavy oils water60-900
gase water 20-300
organic solvent brine 150-500
water brine 600-1200
Gases Brine 15-250
Heaters
Steam Water 1500-4000
Steam organic solvent 500-1000
Steam light oils 300-900
Steam heavy oils 60-450
Steam gass 30-300
Dowtherm Heavy oils 50-300
Dowtherm Gases 20-200
flue gases steam 30-100
flue gases hydrocarbon vapor 30-100
Condensers
Aqueous vapor water 1000-1500
Organic vapor Water 700-1000
Organic (some non condensable gases) Water 500-700
Vacuum condensers Water 200-500
Vaporizers
Steam Aqueuos solutions 1000-1500
Steam Light organics 900-1200
Steam Heavy organics 600-900
Alternatively the overall heat transfer coefficient is evalauted from the individual heat
transfer coefficient as:
1
Uo
=
1
ho
+
1
hod
+
do ln (do/di)
2kw
+
do
di
1
hi
+
do
di
1
hid
(4.11)

4.2 ε- NTU 61
where
Uo = the overall coefficient based on the outside area of the tubeW/m2 o
C
ho = outside fluid film coefficient, W/m2 o
C
hi = inside fluid film coefficient, W/m2 o
C
hod = outside dirt coefficient (Fouling factor), W/m2 o
C
hi = inside dirt coefficient,W/m2 o
C
kw = thermal conductivity of the tube wall material, W/mo
C
do = tube outside diameter, m
di = tube inside diameter, m
4.1.4 Heat transfer coefficient
The heat transfer coefficient is governed by general function for forced convective as
Nu =
hd
k
= f Re, Pr,
d
L
,
µ
µw
(4.12)
and for natural convection as
Nu =
hd
k
= f Gr, Pr,
µ
µw
(4.13)
Design equations for the heat transfer coefficient for various flow geometry (tube, plate)
and configuration are given in Appendix 1. Design equation for the heat transfer coefficient
for condensation and boiling is given also in appendix A.
4.1.5 Fouling factor (hid, hod)
Heat transfer may be degraded in time by corrosion, deposits of reaction products, or-
ganic growths, etc. These effects are accounted for quantitatively by fouling resistances.
Extensive data on fouling factor are given TEMA standards. Typical fouling factors for
common process and service fluids are given in the table 4.2. These values are for shell
and tube heat exchangers with plain (not finned) tubes.
4.2 ε- NTU
The effectiveness (ε) of a heat exchanger is defined as the ratio between the actual heat
load to the maximum possible heat load.
ε =
Q
Qmax
(4.14)
This is related to the heat exchanger size and capacity as
ε = f(NTU, C) (4.15)
Where NTU is number of transfer unit and is defined as
NTU = N =
UA
Cmin
(4.16)
and C is the heat capacity ratio defined using energy equation as:
Q = MhCph(Thi − Tho) = McCpc(Tco − Tci) (4.17)

Table 4.2. Fouling factor
Fluid Coefficient (W/m2 o
C) Factor (resistance (m2 o
C/W)
River water 3000-12000 0.003-0.0001
Sea water 1000-3000 0.001-0.0003
cooling water (towers) 3000-6000 0.0003-0.00017
Towns water (soft) 3000-5000 0.0003-0.0002
Towns water (hard) 1000-2000 0.001-0.0005
Steam condensate 1500-5000 0.00067-0.0002
Steam oil free 4000-10000 0.0025-0.00001
Steam oil traces 2000-5000 0.0005-0.0002
Refrigerated brine 3000-5000 0.0003-0.0002
Air and industrial gases 5000-10000 0.0002-0.00001
Flue gases 2000-5000 0.0005-0.0002
Organic vapor 5000 0.0002
Organic liquids 5000 0.0002
Light hydrocarbons 5000 0.0002
Heavy hydrocarbons 2000 0.0005
Boiling organics 2500 0.0004
Condensing organics 5000 0.0002
Heavy transfer fluids 5000 0.0002
Aqueous salt solutions 3000-5000 0.0003-0.0002
MhCph < McCpc ⇒ Cmin = MhCph, Cmax = McCpc (4.18)
MhCpc > McCpc ⇒ Cmin = McCpc, Cmax = MhCph (4.19)
Qmax = Cmin(Thi − Tci) (4.20)
C =
Cmin
Cmax
(4.21)
εh =
Thi − Tho
Thi − Tci
, εc =
Tco − Tci
Thi − Tci
(4.22)
ε =
∆Tc
Tspan
(4.23)
where Tspan is defined in fig. 4.5 for counter current flow

4.2 ε- NTU 63
Tci
Tco
Thi
ThoTspan
∆θ
0 A
Figure 4.5. Temperature distribution in counter current flow
The ε equation for various heat exchanger configuration is given as
• Parallel flow
ε =
1 − exp [−N(1 + C)]
1 + C
(4.24)
• Counter current flow
ε =
1 − exp [−N(1 + C)]
1 − C exp [−N(1 − C)]
(4.25)
• Cross flow
1. Both fluid unmixed mixed
ε = 1 − exp
exp(−NCn) − 1
Cn
(4.26)
where
n = N−0.22
(4.27)
2. Both fluid mixed
ε =
1
1 − exp(−N) − 1
+
C
1 − exp(−NC) − 1
−
1
N
−1
(4.28)
3. Cmax mixed, Cmin unmixed
ε =
1
C
{1 − exp [−C (1 − exp(−N))]} (4.29)

4. Cmax unmixed, Cmin mixed
ε = 1 − exp −
1
C
[1 − exp(−NC)] (4.30)
• One shell pass, 2,4, 6 tube passes
ε = 2



1 + C + (1 + C2)
1 + exp −N (1 + C2)
1 − exp −N (1 + C2)



−1
(4.31)
• Condenser
ε = 1 − e−N
(4.32)
• Evaporator
ε = 1 − e−N
(4.33)
Alternatively these equations are presented in a graphical form. The various curves of ε
vs NTU can be found in textbooks like Kern (1964( and Perry and Green (2000).
4.3 Link between LMTD and NTU
• Cocurrent
ln
∆T1
∆T2
= ln
Thi − Tci
Tho − Tco
= Nh + Nc (4.34)
• Counter current
ln
∆T1
∆T2
= ln
Thi − Tco
Tho − Tci
= Nh − Nc (4.35)
4.4 The Theta Method
Alternative method of representing the performance of heat exchangers may be given by
Theta method [146] as
Θ =
∆Tm
Tspan
(4.36)
where ∆Tm is the mean temperature difference and Tspan is the maximum temperature
difference (Thi−Tci) (see Fig. 4.5). The Theta method is related is related to the associated
ε and NTU methods by expressions
Θ =
∆Tm
Tspan
=
ε
NTU
(4.37)
The relationship between parameters are often presented in graphical form as shown in
Fig.4.6. However, they all depend on finding ∆Tm or ∆Tlm

4.4 The Theta Method 65
Figure 4.6. θ correction charts for mean temperature diﬀerence: (a) One shell pass and any
multiple of two tube passes. (b) Two shell passes and any multiple of four tube passes.[121].

66 5 Thermal Design
5 Thermal Design
5.1 Design Consideration
5.1.1 Fluid Stream Allocations
There are a number of practical guidelines which can lead to the optimum design of a
given heat exchanger. Remembering that the primary duty is to perform its thermal duty
with the lowest cost yet provide excellent in service reliability, the selection of fluid stream
allocations should be of primary concern to the designer. There are many trade-offs in
fluid allocation in heat transfer coefficients, available pressure drop, fouling tendencies
and operating pressure.
• The higher pressure fluid normally flows through the tube side. With their small
diameter and nominal wall thicknesses, they are easily able to accept high pressures
and avoids more expensive, larger diameter components to be designed for high
pressure. If it is necessary to put the higher pressure stream in the shell, it should
be placed in a smaller diameter and longer shell.
• Place corrosive fluids in the tubes, other items being equal. Corrosion is resisted
by using special alloys and it is much less expensive than using special alloy shell
materials. Other tube side materials can be clad with corrosion resistant materials
or epoxy coated.
• Flow the higher fouling fluids through the tubes. Tubes are easier to clean using
common mechanical methods.
• Because of the wide variety of designs and configurations available for the shell
circuits, such as tube pitch, baffle use and spacing, multiple nozzles, it is best to
place fluids requiring low pressure drops in the shell circuit.
• The fluid with the lower heat transfer coefficient normally goes in the shell circuit.
This allows the use of low-fin tubing to offset the low transfer rate by providing
increased available surface.
Quiz: The top product of a distillation column is condensed using sea water. Allocate
the fluids in the tube and the shell of the heat exchanger?.
5.1.2 Shell and tube velocity
High velocities will give high heat transfer coefficients but also a high pressure drop and
cause erosion. The velocity must be high enough to prevent any suspended solids settling,
but not so high as to cause corrosion. High velocities will reduce fouling. Plastic inserts
are sometimes used to reduce erosion at the tube inlet. Typical design velocity are given
below:
Liquids
1. Tube-side process fluids:1 to 2 m/s, maximum 4 m/s if required to reduce fouling:
water 1.5 to 2.5 m/s
2. Shell side: 0.3 to 1/m/s

5.1 Design Consideration 67
Vapors
For vapors, the velocity used will depend on the operating pressure and fluid density; the
lower values in the range given below will apply to molecular weight materials
Vacuum 50 to 70 m/s
Atmospheric pressure 10 to 30 m/s
High pressure 5 to 10 m/s
5.1.3 Stream temperature
The closer the temperature approach used (the difference between the outlet temperature
of one stream and the inlet temperature of the other stream) the larger will be the heat
transfer area required for a given duty. The optimum value will depend on the application
and can only be determined by making an economic analysis of alternative designs. As
a general guide the greater temperature difference should be at least 20 o
C. and the
least temperature difference 5 to 7 o
C for cooler using cooling water and 3 to 5 o
C using
refrigerated brine. The maximum temperature rise in recirculated cooling water is limited
to around 30 o
C. Care should be taken to ensure that cooling media temperatures are kept
well above the freezing point of the process materials. When heat exchange is between
process fluids for heat recovery the optimum approach temperatures will normally not be
lower than 20 o
C.
5.1.4 Pressure drop
The value suggested below can be used as a general guide and will normally give designs
that are near the optimum.
Liquids
Viscosity<1 mN s/m2
∆p< 35kN/m2
Viscosity=1 to 10mN s/m2
∆p= 50-70 kN/m2
Gas and Vapors
High vacuum 0.4-0.8 kN/m2
Medium vacuum 0.1×absolute pressure
1 to 2 bar 0.5×system gauge pressure
Above 10 bar 0.1×system gauge pressure
When a high-pressure drop is utilized, care must be taken to ensure that the resulting
high fluid velocity does not cause erosion or flow -induced tube vibration.
5.1.5 Fluid physical properties
In the correlation used to predict heat-transfer coefficients, the physical properties are
usually evaluated at the mean stream temperature. This is satisfactory when the tem-
perature change is small, but can cause a significant error when change in temperature
is large. In these circumstances , a simple and safe procedure is to evaluate the heat
transfer coefficients at the stream inlet and outlet temperatures and use the lowest of the

68 5 Thermal Design
two value. Alternatively, the method suggested by Frank (1978) can be used; in which
Q =
A [U2(T1 − t2) − U2(T2 − t1)]
ln U2(T1−t2)
U1(T2−t1)
(5.1)
where U1, U2 are evaluated at the end of the exchanger.
If the variation is too large for these simple methods to be used it will be necessary
to divide the temperature-enthalpy profile into sections and evaluate the heat transfer
coefficients and area required for each section.
5.2 Design data
Before discussing actual thermal design, let us look at the data that must be furnished
by the process licensor before design can begin:
1. flow rates of both streams.
2. inlet and outlet temperatures of both streams.
3. operating pressure of both streams. This is required for gases, especially if the gas
density is not furnished; it is not really necessary for liquids, as their properties do
not vary with pressure.
4. allowable pressure drop for both streams. This is a very important parameter for
heat exchanger design. Generally, for liquids, a value of 0.5-0.7 kg/cm2
is permitted
per shell. A higher pressure drop is usually warranted for viscous liquids, especially
in the tubeside. For gases, the allowed value is generally 0.05-0.2 kg/cm2
, with 0.1
kg/cm2
being typical.
5. fouling resistance for both streams. If this is not furnished, the designer should
adopt values specified in the TEMA standards or based on past experience.
6. physical properties of both streams. These include viscosity, thermal conductivity,
density, and specific heat, preferably at both inlet and outlet temperatures. Viscos-
ity data must be supplied at inlet and outlet temperatures, especially for liquids,
since the variation with temperature may be considerable and is irregular (neither
linear nor log-log).
7. heat duty. The duty specified should be consistent for both the shellside and the
tubeside.
8. type of heat exchanger. If not furnished, the designer can choose this based upon
the characteristics of the various types of construction described earlier. In fact, the
designer is normally in a better position than the process engineer to do this.
9. line sizes. It is desirable to match nozzle sizes with line sizes to avoid expanders
or reducers. However, sizing criteria for nozzles are usually more stringent than for
lines, especially for the shellside inlet. Consequently, nozzle sizes must sometimes be
one size (or even more in exceptional circumstances) larger than the corresponding
line sizes, especially for small lines.
10. preferred tube size. Tube size is designated as O.D., thickness, length. Some plant
owners have a preferred O.D., thickness (usually based upon inventory considera-
tions), and the available plot area will determine the maximum tube length. Many
plant owners prefer to standardize all three dimensions, again based upon inventory
considerations.

5.3 Tubeside design 69
11. maximum shell diameter. This is based upon tube-bundle removal requirements
and is limited by crane capacities. Such limitations apply only to exchangers with
removable tube bundles, namely U-tube and floating-head. For fixed-tubesheet
exchangers, the only limitation is the manufa’s fabrication capability and the avail-
ability of components such as dished ends and flanges. Thus, floating-head heat
exchangers are often limited to a shell I.D. of 1.4-1.5 m and a tube length of 6 m
or 9 m, whereas fixedtubesheet heat exchangers can have shells as large as 3 m and
tubes lengths up to 12 m or more.
12. materials of construction. If the tubes and shell are made of identical materials, all
components should be of this material. Thus, only the shell and tube materials of
construction need to be specified. However, if the shell and tubes are of different
metallurgy, the materials of all principal components should be specified to avoid
any ambiguity. The principal components are shell (and shell cover), tubes, channel
(and channel cover), tubesheets, and baffles. Tubesheets may be lined or clad.
13. special considerations. These include cycling, upset conditions, alternative operating
scenarios, and whether operation is continuous or intermittent.
5.3 Tubeside design
Tubeside calculations are quite straightforward, since tubeside flow represents a simple
case of flow through a circular conduit. Heat-transfer coefficient and pressure drop both
vary with tubeside velocity, the latter more strongly so. A good design will make the best
use of the allowable pressure drop, as this will yield the highest heat-transfer coefficient.
If all the tubeside fluid were to flow through all the tubes (one tube pass), it would lead
to a certain velocity. Usually, this velocity is unacceptably low and therefore has to be
increased. By incorporating pass partition plates (with appropriate gasketing) in the
channels, the tubeside fluid is made to flow several times through a fraction of the total
number of tubes. Thus, in a heat exchanger with 200 tubes and two passes, the fluid flows
through 100 tubes at a time, and the velocity will be twice what it would be if there were
only one pass. The number of tube passes is usually one, two, four, six, eight, and so on.
5.3.1 Heat-transfer coefficient
The tubeside heat-transfer coefficient is a function of the Reynolds number, the Prandtl
number, and the tube diameter. These can be broken down into the following fundamen-
tal parameters: physical properties (namely viscosity, thermal conductivity, and specific
heat); tube diameter; and, very importantly, mass velocity.
The variation in liquid viscosity is quite considerable; so, this physical property has the
most dramatic effect on heat-transfer coefficient. The fundamental equation for turbulent
heat-transfer inside tubes is:
Nu = CRea
Prb µ
µw
c
, (5.2)
or
h = C
k
D
GD
µ
a
Cpµ
k
b
µ
µw
c
(5.3)

70 5 Thermal Design
where
Nu = hde
k
Nusselt number
Pr = Cpµ
k
Prandtl number
Re ρud
µ
Reynolds number
de
4A
P
hydraulic diameter
A cross-sectional area
P wetted perimeter
u fluid velocity
µw fluid viscosity at the tube wall temperature
k fluid thermal conductivity
Cp fluid specific heat
C =



0.021 gases
0.023 non-viscous liquid
0.027 viscous liquid
a = 0.8
b = 0.3 for cooling
b = 0.4 for heating
c = 0.14
Viscosity influences the heat-transfer coefficient in two opposing ways- as a parameter of
the Reynolds number, and as a parameter of Prandtl number. Thus, from Eq. 5.3:
h ∝ µ0.8−0.33
= µ0.47
(5.4)
In other words, the heat-transfer coefficient is inversely proportional to viscosity to the
0.47 power. Similarly, the heat-transfer coefficient is directly proportional to thermal
conductivity to the 0.67 power.
These two facts lead to some interesting generalities about heat transfer. A high thermal
conductivity promotes a high heat-transfer coefficient. Thus, cooling water (thermal
conductivity of around 0.55 kcal/hm◦
C) has an extremely high heat-transfer coefficient
of typically 6,000 kcal/hm2◦
C, followed by hydrocarbon liquids (thermal conductivity
between 0.08 and 0.12 kcal/hm◦
C) at 250-1,300 kcal/hm2◦
C, and then hydrocarbon gases
(thermal conductivity between 0.02 and 0.03 kcal/hm◦
C) at 50-500 kcal/hm2◦
C.
Hydrogen is an unusual gas, because it has an exceptionally high thermal conductivity
(greater than that of hydrocarbon liquids). Thus, its heat-transfer coefficient is toward
the upper limit of the range for hydrocarbon liquids.
The range of heat-transfer coefficients for hydrocarbon liquids is rather large due to the
large variation in their viscosity, from less than 0.1 cP for ethylene and propylene to more
than 1,000 cP or more for bitumen. The large variation in the heat-transfer coefficients
of hydrocarbon gases is attributable to the large variation in operating pressure. As
operating pressure rises, gas density increases. Pressure drop is directly proportional to
the square of mass velocity and inversely proportional to density. Therefore, for the same
pressure drop, a higher mass velocity can be maintained when the density is higher. This
larger mass velocity translates into a higher heat-transfer coefficient.
5.3.2 Pressure drop
The pressure drop due to friction exists because of the shear stress between the fluid and
the tube wall. Estimation of the friction pressure drop is somewhat more complex and

5.3 Tubeside design 71
various approaches have been taken, for example the frictional pressure gradient is given
as
−
dp
dz f
=
4τo
d
=
4fG2
2dρ
, (5.5)
where G is the mass flux in kg/m2
s and f is the friction factor calculated using a Blasius-
type model as
f =



0.3164
Re0.25 Re ≥ 2320
64
Re
Re < 2320 .
Integration of equation B.1 yields
∆p =
4fG2
2ρ
L
d
, (5.6)
Mass velocity strongly influences the heat-transfer coefficient. For turbulent flow, the
tubeside heat-transfer coefficient varies to the 0.8 power of tubeside mass velocity, whereas
tubeside pressure drop varies to the square of mass velocity. Thus, with increasing mass
velocity, pressure drop increases more rapidly than does the heat-transfer coefficient.
Consequently, there will be an optimum mass velocity above which it will be wasteful to
increase mass velocity further.
Furthermore, very high velocities lead to erosion. However, the pressure drop limitation
usually becomes controlling long before erosive velocities are attained. The minimum
recommended liquid velocity inside tubes is 1.0 m/s, while the maximum is 2.5-3.0 m/s.
Pressure drop is proportional to the square of velocity and the total length of travel.
Thus, when the number of tube passes is increased for a given number of tubes and a
given tubeside flow rate, the pressure drop rises to the cube of this increase. In actual
practice, the rise is somewhat less because of lower friction factors at higher Reynolds
numbers, so the exponent should be approximately 2.8 instead of 3.
Tubeside pressure drop rises steeply with an increase in the number of tube passes. Con-
sequently, it often happens that for a given number of tubes and two passes, the pressure
drop is much lower than the allowable value, but with four passes it exceeds the allowable
pressure drop. If in such circumstances a standard tube has to be employed, the designer
may be forced to accept a rather low velocity. However, if the tube diameter and length
may be varied, the allowable pressure drop can be better utilized and a higher tubeside
velocity realized.
The following tube diameters are usually used in the CPI: (1/4, 3/8, 1/2, 5/8, 3/4, 1, 1
1/4, 1 1/2 in. Of these, 3/4 in. and 1 in. are the most popular. Tubes smaller than 3/4
in. O.D. should not be used for fouling services. The use of small-diameter tubes, such as
1 in., is warranted only for small heat exchangers with heat-transfer areas less than 20-30
m2
.
It is important to realize that the total pressure drop for a given stream must be met.
The distribution of pressure drop in the various heat exchangers for a given stream in a
particular circuit may be varied to obtain good heat transfer in all the heat exchangers.
Consider a hot liquid stream flowing through several preheat exchangers. Normally, a
pressure drop of 0.7 kg/cm2
per shell is permitted for liquid streams. If there are five
such preheat exchangers, a total pressure drop of 3.5 kg/cm2
for the circuit would be
permitted. If the pressure drop through two of these exchangers turns out to be only 0.8
kg/cm2
, the balance of 2.7 kg/cm2
would be available for the other three.

72 5 Thermal Design
5.4 Shell side design
Shell side design The shellside calculations are far more complex than those for the tube-
side. This is mainly because on the shellside there is not just one flow stream but one
principal cross-flow stream and four leakage or bypass streams. There are various shell-
side flow arrangements, as well as various tube layout patterns and baffling designs, which
together determine the shellside stream analysis.
5.4.1 Shell configuration
TEMA defines various shell patterns based on the flow of the shellside fluid through the
shell: E, F, G, H, J, K, and X (see Figure 3.1).
In a TEMA E single-pass shell, the shellside fluid enters the shell at one end and leaves
from the other end. This is the most common shell type - more heat exchangers are built
to this configuration than all other configurations combined.
A TEMA F two-pass shell has a longitudinal baffle that divides the shell into two passes.
The shellside fluid enters at one end, traverses the entire length of the exchanger through
one-half the shell cross-sectional area, turns around and flows through the second pass,
then finally leaves at the end of the second pass. The longitudinal baffle stops well short
of the tubesheet, so that the fluid can flow into the second pass.
The F shell is used for temperature- cross situations - that is, where the cold stream leaves
at a temperature higher than the outlet temperature of the hot stream. If a two-pass (F)
shell has only two tube passes, this becomes a true countercurrent arrangement where a
large temperature cross can be achieved.
A TEMA J shell is a divided-flow shell wherein the shellside fluid enters the shell at the
center and divides into two halves, one flowing to the left and the other to the right and
leaving separately. They are then combined into a single stream. This is identified as a
J 1-2 shell. Alternatively, the stream may be split into two halves that enter the shell at
the two ends, flow toward the center, and leave as a single stream, which is identified as
a J 2-1 shell.
A TEMA G shell is a split-flow shell (see Figure 3.1). This construction is usually em-
ployed for horizontal thermosyphon reboilers. There is only a central support plate and
no baffles. A G shell cannot be used for heat exchangers with tube lengths greater than
3 m, since this would exceed the limit on maximum unsupported tube length specified by
TEMA - typically 1.5 m, though it varies with tube O.D., thickness, and material.
When a larger tube length is needed, a TEMA H shell (see Figure3.1) is used. An H shell
is basically two G shells placed side-by-side, so that there are two full support plates. This
is described as a double-split configuration, as the flow is split twice and recombined twice.
This construction, too, is invariably employed for horizontal thermosyphon reboilers. The
advantage of G and H shells is that the pressure drop is drastically less and there are no
cross baffles.
A TEMA X shell (see Figure 3.1) is a pure cross-flow shell where the shellside fluid enters
at the top (or bottom) of the shell, flows across the tubes, and exits from the opposite side
of the shell. The flow may be introduced through multiple nozzles located strategically
along the length of the shell in order to achieve a better distribution. The pressure drop
will be extremely low - in fact, there is hardly any pressure drop in the shell, and what
pressure drop there is, is virtually all in the nozzles. Thus, this configuration is employed
for cooling or condensing vapors at low pressure, particularly vacuum. Full support plates
can be located if needed for structural integrity; they do not interfere with the shellside
flow because they are parallel to the flow direction.
A TEMA K shell (see Figure 3.1) is a special cross-flow shell employed for kettle reboilers
(thus the K). It has an integral vapor-disengagement space embodied in an enlarged shell.
Here, too, full support plates can be employed as required.

5.4 Shell side design 73
5.4.2 Tube layout patterns
There are four tube layout patterns, as shown in Figure 5.1: triangular (30◦
), rotated
triangular (60◦
), square (90◦
), and rotated square (45◦
).
Figure 5.1. Tubes layout pattern.
A triangular (or rotated triangular) pattern will accommodate more tubes than a square
(or rotated square) pattern. Furthermore, a triangular pattern produces high turbulence
and therefore a high heat-transfer coefficient. However, at the typical tube pitch of 1.25
times the tube O.D., it does not permit mechanical cleaning of tubes, since access lanes
are not available. Consequently, a triangular layout is limited to clean shellside services.
For services that require mechanical cleaning on the shellside, square patterns must be
used. Chemical cleaning does not require access lanes, so a triangular layout may be used
for dirty shellside services provided chemical cleaning is suitable and effective.
A rotated triangular pattern seldom offers any advantages over a triangular pattern, and
its use is consequently not very popular.
For dirty shellside services, a square layout is typically employed. However, since this is an
in-line pattern, it produces lower turbulence. Thus, when the shellside Reynolds number
is low (< 2,000), it is usually advantageous to employ a rotated square pattern because
this produces much higher turbulence, which results in a higher efficiency of conversion
of pressure drop to heat transfer.
As noted earlier, fixed-tubesheet construction is usually employed for clean services on
the shellside, Utube construction for clean services on the tubeside, and floating-head
construction for dirty services on both the shellside and tubeside. (For clean services
on both shellside and tubeside, either fixed-tubesheet or U-tube construction may be
used, although U-tube is preferable since it permits differential expansion between the
shell and the tubes.) Hence, a triangular tube pattern may be used for fixed-tubesheet
exchangers and a square (or rotated square) pattern for floating-head exchangers. For
U-tube exchangers, a triangular pattern may be used provided the shellside stream is
clean and a square (or rotated square) pattern if it is dirty.
5.4.3 Tube pitch
Tube pitch is defined as the shortest distance between two adjacent tubes.

74 5 Thermal Design
For a triangular pattern, TEMA specifies a minimum tube pitch of 1.25 times the tube
O.D. Thus, a 25- mm tube pitch is usually employed for 20-mm O.D. tubes.
For square patterns, TEMA additionally recommends a minimum cleaning lane of 4 in.
(or 6 mm) between adjacent tubes. Thus, the minimum tube pitch for square patterns
is either 1.25 times the tube O.D. or the tube O.D. plus 6 mm, whichever is larger. For
example, 20-mm tubes should be laid on a 26-mm (20 mm + 6 mm) square pitch, but
25-mm tubes should be laid on a 31.25-mm (25 mm ´ 1.25) square pitch.
Designers prefer to employ the minimum recommended tube pitch, because it leads to
the smallest shell diameter for a given number of tubes. However, in exceptional cir-
cumstances, the tube pitch may be increased to a higher value, for example, to reduce
shellside pressure drop. This is particularly true in the case of a cross-flow shell.
5.4.4 Baffling
Type of baffles. Baffles are used to support tubes, enable a desirable velocity to be
maintained for the shellside fluid, and prevent failure of tubes due to flow-induced vibra-
tion. There are two types of baffles: plate and rod. Plate baffles may be single-segmental,
double-segmental, or triple-segmental, as shown in Figure 5.2.
Figure 5.2. Types of baffles.
Baffle spacing. Baffle spacing is the centerline-to-centerline distance between adjacent
baffles. It is the most vital parameter in STHE design.
The TEMA standards specify the minimum baffle spacing as one-fifth of the shell inside
diameter or 2 in., whichever is greater. Closer spacing will result in poor bundle pene-
tration by the shellside fluid and difficulty in mechanically cleaning the outsides of the
tubes. Furthermore, a low baffle spacing results in a poor stream distribution as will be
explained later.
The maximum baffle spacing is the shell inside diameter. Higher baffle spacing will
lead to predominantly longitudinal flow, which is less efficient than cross-flow, and large
unsupported tube spans, which will make the exchanger prone to tube failure due to
flow-induced vibration.
Optimum baffle spacing. For turbulent flow on the shellside (Re > 1,000), the heat-
transfer coefficient varies to the 0.6-0.7 power of velocity; however, pressure drop varies

to the 1.7-2.0 power. For laminar flow (Re < 100), the exponents are 0.33 for the heat-
transfer coefficient and 1.0 for pressure drop. Thus, as baffle spacing is reduced, pressure
drop increases at a much faster rate than does the heat-transfer coefficient.
This means that there will be an optimum ratio of baffle spacing to shell inside diameter
that will result in the highest efficiency of conversion of pressure drop to heat transfer.
This optimum ratio is normally between 0.3 and 0.6.
Baffle cut. As shown in Figure 5.3, baffle cut is the height of the segment that is cut in
each baffle to permit the shellside fluid to flow across the baffle. This is expressed as a
percentage of the shell inside diameter. Although this, too, is an important parameter
for STHE design, its effect is less profound than that of baffle spacing.
Figure 5.3. Baffle cut.
Baffle cut can vary between 15% and 45% of the shell inside diameter.
Both very small and very large baffle cuts are detrimental to efficient heat transfer on the
shellside due to large deviation from an ideal situation, as illustrated in Figure 5.4.
Figure 5.4. Effect of small and large baffle cuts.
It is strongly recommended that only baffle cuts between 20% and 35% be employed. Re-
ducing baffle cut below 20% to increase the shellside heat-transfer coefficient or increasing
the baffle cut beyond 35% to decrease the shellside pressure drop usually lead to poor de-
signs. Other aspects of tube bundle geometry should be changed instead to achieve those
goals. For example, doublesegmental baffles or a divided-flow shell, or even a cross-flow
shell, may be used to reduce the shellside pressure drop.

76 5 Thermal Design
For single-phase fluids on the shellside, a horizontal baffle cut (Figure 5.5) is recommended,
because this minimizes accumulation of deposits at the bottom of the shell and also
prevents stratification. However, in the case of a two-pass shell (TEMA F), a vertical cut
is preferred for ease of fabrication and bundle assembly.
Figure 5.5. Baffle cut orientation
5.4.5 Equalize cross-flow and window velocities
Flow across tubes is referred to as cross-flow, whereas flow through the window area (that
is, through the baffle cut area) is referred to as window flow.
The window velocity and the cross-flow velocity should be as close as possible - preferably
within 20%
of each other. If they differ by more than that, repeated acceleration and deceleration take
place along the length of the tube bundle, resulting in inefficient conversion of pressure
drop to heat transfer.
5.4.6 Shellside stream analysis (Flow pattern)
On the shellside, there is not just one stream, but a main cross-flow stream and four
leakage or bypass streams, as illustrated in Figure 5.6. Tinker (4) proposed calling these
streams the main cross-flow stream (B), a tube-to-baffle-hole leakage stream (A), a bundle
bypass stream (C), a pass-partition bypass stream (F), and a baffle-to-shell leakage stream
(E). While the B (main cross-flow) stream is highly effective for heat transfer, the other
streams are not as effective. The A stream is fairly efficient, because the shellside fluid
is in contact with the tubes. Similarly, the C stream is in contact with the peripheral
tubes around the bundle, and the F stream is in contact with the tubes along the pass-
partition lanes. Consequently, these streams also experience heat transfer, although at
a lower efficiency than the B stream. However, since the E stream flows along the shell
wall, where there are no tubes, it encounters no heat transfer at all.
The fractions of the total flow represented by these five streams can be determined for a
particular set of exchanger geometry and shellside flow conditions by any sophisticated
heatexchanger thermal design software. Essentially, the five streams are in parallel and
flow along paths of varying hydraulic resistances. Thus, the flow fractions will be such that
the pressure drop of each stream is identical, since all the streams begin and end at the
inlet and outlet nozzles. Subsequently, based upon the efficiency of each of these streams,
the overall shellside stream efficiency and thus the shellside heat-transfer coefficient is
established.

Figure 5.6. Tube arrangement
Since the flow fractions depend strongly upon the path resistances, varying any of the
following construction parameters will affect stream analysis and thereby the shellside
performance of an exchanger:
• baffle spacing and baffle cut;
• tube layout angle and tube pitch;
• number of lanes in the flow direction and lane width;
• clearance between the tube and the baffle hole;
• clearance between the shell I.D. and the baffle; and
• location of sealing strips and sealing rods.
Using a very low baffle spacing tends to increase the leakage and bypass streams. This
is because all five shellside streams are in parallel and, therefore, have the same pressure
drop. The leakage path dimensions are fixed. Consequently, when baffle spacing is de-
creased, the resistance of the main cross-flow path and thereby its pressure drop increases.
Since the pressure drops of all five streams must be equal, the leakage and bypass streams
increase until the pressure drops of all the streams balance out. The net result is a rise
in the pressure drop without a corresponding increase in the heat-transfer coefficient.
The shellside fluid viscosity also affects stream analysis profoundly. In addition to influ-
encing the shellside heat transfer and pressure drop performance, the stream analysis also
affects the mean temperature difference (MTD) of the exchanger. This will be discussed
in detail later. First, though, let’s look at an example that demonstrates how to optimize
baffle design when there is no significant temperature profile distortion.
5.4.7 Heat transfer coefficient and pressure drop
For the shell side heat transfer coefficient and pressure drop there are a number of methods
these include:
• Kern’s method
• Donohue’s method
• Bell-Delaware method
• Tinker’s method

78 5 Thermal Design
Besides these methods there is some proprietary methods putout by various organization
for use by their member companies. A number of these method are based on one of the
above methods. Some are based upon a judicious combination of methods 3 and 4 above
and supplemented by further research data. Among the most popular of the proprietary
methods, judged by their large clientele are
• Heat Transfer Research Inc. (HTRI), Alliambra, california. This method is also
known as stream analysis method.
• Heat Transfer and Fluid Flow Service (HTFS), Engineering Science Division, AERE,
Harwell, United Kingdom Method.
In this work only Kern’s method is given below. Bell-Delaware method may be found in
Coulson and Richardson’s
5.4.8 Heat transfer coefficient
Nu = 0.36Re0.55
Pr1/3 µ
µw
0.14
, (5.7)
where
Nu = hde
k
Nusselt number
Pr = Cpµ
k
Prandtl number
Re = Gde
µ
Reynolds number
de = 4A
P
hydraulic diameter
A = cross-sectional flow area
P = wetted perimeter
G = M
As
Mass flux
As = (pt−do)DslB
pt
fluid viscosity at the tube wall temperature
pt = pitch diameter
Ds = shell diameter
lB = Baffle spacing
Hydraulic diameter (Fig. 5.1)
de =



p2
t −πd2
o/4
πdo
for square pitch
0.87p2
t /2−πd2
o/8
πdo/2
for equilateral triangular pitch
5.4.9 Pressure drop
∆p = 4f
Ds
d
ρu2
2
L
lb
µ
µw
−0.14
, (5.8)
where
f =



0.3164
Re0.25 Re ≥ 2320
64
Re
Re < 2320 .
L=tube length
lB = baffle spacing. The term (L/lB) is the number of times the flow crosses the tube
bundle=(NB + 1). Where NB is the number of baffles.

5.5 Design Algorithm 79
5.5 Design Algorithm
Step1
Specification
Define duty Q
Make energy balance if needed
to calcualted unspecified flow
rates or temperature
Q=M )c pc c2 c1 h ph h1 h2c (T -T )=M C (T -T
Step2
Calculate physical properties
Step3
Assume value of overall
coefficient Uo,ass
Step 4
Decide number of shell and
tube passes
Calculate T∆ ∆T , F andlm m
Step 5
Determine heat transfer area
required A To o,ass=q/U ∆ m
Step 6
Decide type, tube size, material,
layout
Assign fluids to shell or tube
Step 7
Calculate number of tubes
Step 8
Calculate shell diameter
Step 9
Estimate tube-side heat
transfer coefficient
Step 10
Decide baffle spacing and estimate
shell side heat transfer coefficient
Step 11
Calculate overall heat transfer
Coefficient including fouling factors
Uo,cal
Step 12
Estimate tube and shell side
pressure drop
Step 13
Estimate cost of heat exchanger
Can design be
optmized to
reduce cost?
Accept design
Is pressure drops
within specification?
0<(Uo,cal o,ass o,ass-U )/U <30Set Uo,ass=Uo,cal
Yes
No
No
Figure 5.7. Design procedure for shell and tube heat exchanger.

80 6 Specification sheet
6 Specification sheet
Specification sheet is a data sheet that contains the information provided by the customer
to the vendor for the process and mechanical designs of an exchanger. After the process
design is done, the engineer fills in some further information. The rest of the information
is filled after the mechanical design is completed. The specification sheet is a medium of
communication between different parties involved in the procurement, design and fabri-
cation of heat exchanger. It is also used to compare the performance of the installed unit
with the design conditions.
6.1 Information included
The information contained in the sheet is best decribed by a data sheet. Although each
company has its own version of data sheet, the most popular one is that of the TEMA
standards. It is similar to that of API standard 660. It contains the fluid
• flow rate and properties,
• heat duty,
• heat transfer coefficient,
• fouling resistance,
• details about the shell and tube size,
• materials,
• baffle nozzle, etc..
Some variations include information for alternate designs and different systems of units
(British, SI, metric).
6.2 Information not included
The regarding the type of flanges, studs, vent and relief valves, drains lines, welding,
inspection and testing requirement of the material of construction, etc.. are not given in
the specification sheet.
6.3 Operation conditions
The following operating conditions regarding the exchanger operation should be known
to the thermal designer for critical application.
1. Start-up condition and procedure
2. Normal operating conditions
3. Upset and emergency conditions
4. shut down conditions and procedure
5. possibility of switching the shell-side and tube tube side fluid for better design
6. possibility of increasing the allowable pressure drop to control the fouling
7. beside these the spec-sheet should provided with other information concerning the
composition of the streams, their thermal and physical properties and any phase
change occurring.

6.4 Bid evaluation 81
6.4 Bid evaluation
6.4.1 Factor to be consider
For ease in evaluations of the bids submitted by competitive bidders, all pertinent data
from each bid should be put on a large data sheet. During evaluation the following factor
should be kept in mid:
1. The design submitted by the bidders should meet the heat transfer and pressure
drop requirements. Set the upper and lower limit of pressure drop for each bid.
2. if the designs offered by bidder vary, the spec-sheet provided to them should be
checked to see if any anomalies exist
3. Adequate vent, drainage and safety valve should be provided
4. Units should not have hot spot or dead zones
5. Information about vibration analysis must be checked
6. for fouling on the shell side, the tube lay out should permit easy cleaning
7. The fabrication shop should have a good reputation and certificate of inspection
8. The material of construction should be available at the country of the bidder or
their import should not pose any difficulty
9. the delivery should be on schedule
10. cost should be low, cost escalation should be included
11. the payment, penalty, and guarantee clauses in the contact should be evenly balance
and be unduly favorable to the bidder

82 6 Speciﬁcation sheet
Figure 6.1. Data sheet

83
7 Storage, Installation, Operation and Maintenance
Proper storage, installation handling and correct start up emergency, and shutdown pro-
cedure are important for the successful working of a well designed and fabricated heat
exchanger. regular cleaning, maintenance and repairs are necessary to ensure trouble free
operation of the unit for its designed life span. These will be discussed in the following
sections.
NOTE: Before placing your equipment in operation, environment and service conditions
should be checked for compatibility with materials of construction. Contact your nearest
heat exchanger Standard representative if you are not sure what the actual materials of
construction are.
Successful performance of heat transfer equipment, length of service and freedom from
operating diﬃculties are largely dependent upon:
1. Proper thermal design.
2. Proper physical design.
3. Storage practice prior to installation.
4. Manner of installation, including design of foundation and piping.
5. The method of operation.
6. The thoroughness and frequency of cleaning.
7. The materials, workmanship, and tools used in maintenance and making repairs
and replacements.
Failure to perform properly may be due to one or more of the following:
1. Exchanger being dirty.
2. Failure to remove preservation materials after storage.
3. Operating conditions being diﬀerent than design conditions.
4. Air or gas binding.
5. Incorrect piping connections.
6. Excessive clearances between internal parts due to corrosion.
7. Improper application.
7.1 Storage
Standard heat exchangers are protected against the elements during shipment. If they
cannot be installed and put into operation immediately upon receipt at the jobsite, cer-
tain precautions are necessary to prevent deterioration during storage. Responsibility for
integrity of the heat exchangers must be assumed by the user. The manufacturer will not
be responsible for damage, corrosion or other deterioration of heat exchanger equipment
during transit and storage.
Good storage practices are important, considering the high costs of repair or replacement,
and the possible delays for items which require long lead times for manufacture. The
following suggested practices are provided solely as a convenience to the user, who shall
make his own decision on whether to use all or any of them.

84 7 Storage, Installation, Operation and Maintenance
1. On receipt of the heat exchanger, inspect for shipping damage to all protective cov-
ers. If damage is evident, inspect for possible contamination and replace protective
covers as required. If damage is extensive, notify the carrier immediately.
2. If the heat exchanger is not to be placed in immediate service, take precautions to
prevent rusting or contamination.
3. Heat exchangers for oil service, made of ferrous materials, may be pressure-tested
with oil at the factory. However, the residual oil coating on the inside surfaces of
the exchanger does not preclude the possibility of rust formation. Upon receipt,
fill these exchangers with appropriate oil or coat them with a corrosion prevention
compound for storage. These heat exchangers have a large warning decal, indicating
that they should be protected with oil.
4. The choice of preservation of interior surfaces during storage for other service appli-
cations depends upon your system requirements and economics. Only when included
in the original purchase order specifications will specific preservation be incorporated
prior to shipment from the factory.
5. Remove any accumulations of dirt, water, ice or snow and wipe dry before moving
exchangers into indoor storage. If unit was not filled with oil or other preservative,
open drain plugs to remove any accumulated moisture, then reseal. Accumulation
of moisture usually indicates rusting has already started and remedial action should
be taken.
6. Store under cover in a heated area, if possible. The ideal storage environment for
heat exchangers and accessories is indoors, above grade, in a dry, low humidity at-
mosphere which is sealed to prevent entry of blowing dust, rain or snow. Maintain
temperatures between 70◦
F and 105◦
F (wide temperature swings may cause con-
densation and ”sweating” of steel parts). Cover windows to prevent temperature
variations caused by sunlight. Provide thermometers and humidity indicators at
several points, and maintain atmosphere at 40% relative humidity or lower.
7. In tropical climates, it may be necessary to use trays of renewable dessicant (such as
silica gel), or portable dehumidifiers, to remove moisture from the air in the storage
enclosure. Thermostatically controlled portable heaters (vented to outdoors) may
be required to maintain even air temperatures inside the enclosure.
8. Inspect heat exchangers and accessories frequently while they are in storage. Start
a log to record results of inspections and maintenance performed while units are
in storage. A typical log entry should include, for each component, at least the
following:
(a) Date
(b) Inspector’s name
(c) Identification of unit or item
(d) Location
(e) Condition of paint or coating
(f) Condition of interior
(g) Is free moisture present?
(h) Has dirt accumulated?
(i) Corrective steps taken
9. To locate ruptured or corroded tubes or leaking joints between tubes and tubesheets,
the following procedure is recommended:

7.2 Installation 85
• Remove tube side channel covers or bonnets.
• Pressurize the shell side of the exchanger with a cold fluid, preferably water.
• Observe tube joints and tube ends for indication of test fluid leakage.
10. With certain styles of exchangers, it will be necessary to buy or make a test ring to
seal off the space between the floating tubesheet and inside shell diameter to apply
the test in paragraph
11. Consult your nearest sales representative for reference drawings showing installation
of a test ring in your heat exchanger.
12. To tighten a leaking tube joint, use a suitable parallel roller tube expander.
• Do not roll tubes beyond the back face of the tubesheet. Maximum rolling
depth should be tubesheet thickness minus 1/8”.
• Do not re-roll tubes that are not leaking since this needlessly thins the tube
wall.
13. It is recommended that when a heat exchanger is dismantled, new gaskets be used
in reassembly.
• Composition gaskets become brittle and dried out in service and do not provide
an effective seal when reused.
• Metal or metal jacketed gaskets in initial compression match the contact sur-
faces and tend to work-harden and cannot be recompressed on reuse.
14. Use of new bolting in conformance with dimension and ASTM specifications of the
original design is recommended where frequent dismantling is encountered. CAU-
TION: Do not remove channel covers, shell covers, floating head covers or bonnets
until all pressure in the heat exchanger has been relieved and both shell side and
tube side are completely drained.
15. If paint deterioration begins, as evidenced by discoloration or light rusting, consider
touch-up or repainting. If the unit is painted with our standard shop enamel, areas
of light rust may be wire brushed and touched-up with any good quality air-drying
synthetic enamel. Units painted with special paints (when specified on customers’
orders) may require special techniques for touch-up or repair. Obtain specific infor-
mation from the paint manufacturer. Painted steel units should never be permitted
to rust or deteriorate to a point where their strength will be impaired. But a light
surface rusting, on steel units which will be re-painted after installation, will not
generally cause any harm. (See Items 3 and 4 for internal surface preservation.)
16. If the internal preservation (Items 3 and 4 ) appears inadequate during storage,
consider additional corrosion prevention measures and more frequent inspections.
Interiors coated with rust preventive should be restored to good condition and re-
coated promptly if signs of rust occur.
7.2 Installation
7.2.1 Installation Planning
1. On removable bundle heat exchangers, provide sufficient clearance at the stationary
end to permit the removal of the tube bundle from the shell. On the floating head
end, provide space to permit removal of the shell cover and floating head cover.
2. On fixed bundle heat exchangers, provide sufficient clearance at one end to permit
removal and replacement of tubes and at the other end provide sufficient clearance
to permit tube rolling.

3. Provide valves and bypasses in the piping system so that both the shell side and
tube side may be bypassed to permit isolation of the heat exchanger for inspection,
cleaning and repairs.
4. Provide convenient means for frequent cleaning as suggested under maintenance.
5. Provide thermometer wells and pressure gauge pipe taps in all piping to and from
the heat exchanger, located as close to the heat exchanger as possible.
6. Provide necessary air vent valves for the heat exchanger so that it can be purged to
prevent or relieve vapor or gas binding on both the tube side and shell side.
7. Provide adequate supports for mounting the heat exchanger so that it will not settle
and cause piping strains. Foundation bolts should be set accurately. In concrete
footings, pipe sleeves at least one pipe size larger than the bolt diameter slipped over
the bolt and cast in place are best for this purpose as they allow the bolt centers to
be adjusted after the foundation has set.
8. Install proper liquid level controls and relief valves and liquid level and temperature
alarms, etc.
9. Install gauge glasses or liquid level alarms in all vapor or gas spaces to indicate any
failure occurring in the condensate drain system and to prevent flooding of the heat
exchanger.
10. Install a surge drum upstream from the heat exchanger to guard against pulsation
of fluids caused by pumps, compressors or other equipment.
11. Do not pipe drain connections to a common closed manifold; it makes it more
difficult to determine that the exchanger has been thoroughly drained.
7.2.2 Installation at Jobsite
1. If you have maintained the heat exchanger in storage, thoroughly inspect it prior to
installation. Make sure it is thoroughly cleaned to remove all preservation materials
unless stored full of the same oil being used in the system, or the coating is soluble
in the lubricating system oil. If the exchanger was oil-tested by any Standard and
your purchase order did not specify otherwise, the oil used was Tectyl 754, a light-
bodied oil which is soluble in most lubricating oils. Where special preservations were
applied, you should consult the preservative manufacturer’s product information
data for removal instructions.
2. If the heat exchanger is not being stored, inspect for shipping damage to all pro-
tective covers upon receipt at the jobsite. If damage is evident, inspect for possible
contamination and replace protective covers as required. If damage is extensive,
notify the carrier immediately.
3. When installing, set heat exchanger level and square so that pipe connections can
be made without forcing.
4. Before piping up, inspect all openings in the heat exchanger for foreign material.
Remove all wooden plugs, bags of dessicant and shipping covers immediately prior to
installing. Do not expose internal passages of the heat exchanger to the atmosphere
since moisture or harmful contaminants may enter the unit and cause severe damage
to the system due to freezing and/or corrosion.
5. After piping is complete, if support cradles or feet are fixed to the heat exchanger,
loosen foundation bolts at one end of the exchanger to allow free movement. Over-
sized holes in support cradles or feet are provided for this purpose.

7.3 Operation 87
6. If heat exchanger shell is equipped with a bellows-type expansion joint, remove
shipping supports per instructions.
7.3 Operation
1. Be sure entire system is clean before starting operation to prevent plugging of tubes
or shell side passages with refuse. The use of strainers or settling tanks in pipelines
leading to the heat exchanger is recommended.
2. Open vent connections before starting up.
3. Start operating gradually. See Table 1 for suggested start-up and shut-down proce-
dures for most applications. If in doubt, consult the nearest manufactuerer repre-
sentative for specific instructions.
4. After the system is completely filled with the operating fluids and all air has been
vented, close all manual vent connections.
5. Re-tighten bolting on all gasketed or packed joints after the heat exchanger has
reached operating temperatures to prevent leaks and gasket failures. Standard pub-
lished torque values do not apply to packed end joints.
6. Do not operate the heat exchanger under pressure and temperature conditions in
excess of those specified on the nameplate.
7. To guard against water hammer, drain condensate from steam heat exchangers and
similar apparatus both when starting up and shutting down.
8. Drain all fluids when shutting down to eliminate possible freezing and corroding.
9. In all installations there should be no pulsation of fluids, since this causes vibration
and will result in reduced operating life.
10. Under no circumstances is the heat exchanger to be operated at a flowrate greater
than that shown on the design specifications. Excessive flows can cause vibration
and severely damage the heat exchanger tube bundle.
11. Heat exchangers that are out of service for extended periods of time should be
protected against corrosion as described in the storage requirements for new heat
exchangers. Heat exchangers that are out of service for short periods and use water
as the flowing medium should be thoroughly drained and blown dry with warm air,
if possible. If this is not practical, the water should be circulated through the heat
exchanger on a daily basis to prevent stagnant water conditions that can ultimately
precipitate corrosion.

1. Clean exchangers subject to fouling (scale, sludge deposits, etc.) periodically, de-
pending on specific conditions. A light sludge or scale coating on either side of the
tube greatly reduces its effectiveness. A marked increase in pressure drop and/or
reduction in performance usually indicates cleaning is necessary. Since the difficulty
of cleaning increases rapidly as the scale thickens or deposits increase, the intervals
between cleanings should not be excessive.
2. Neglecting to keep tubes clean may result in random tube plugging. Consequent
overheating or cooling of the plugged tubes, as compared to surrounding tubes, will
cause physical damage and leaking tubes due to differential thermal expansion of
the metals.

7.3 Operation 89
3. To clean or inspect the inside of the tubes, remove only the necessary tube side
channel covers or bonnets, depending on type of exchanger construction.
4. If the heat exchanger is equipped with sacrificial anodes or plates, replace these as
required.
5. To clean or inspect the outside of the tubes, it may be necessary to remove the tube
bundle. (Fixed tubesheet exchanger bundles are non-removable).
6. When removing tube bundles from heat exchangers for inspection or cleaning, ex-
ercise care to see that they are not damaged by improper handling.
• The weight of the tube bundle should not be supported on individual tubes
but should be carried by the tubesheets, support or baffle plates or on blocks
contoured to the periphery of the tube bundles.
• Do not handle tube bundles with hooks or other tools which might damage
tubes. Move tube bundles on cradles or skids.
• To withdraw tube bundles, pass rods through two or more of the tubes and
take the load on the floating tubesheet.
• Rods should be threaded at both ends, provided with nuts, and should pass
through a steel bearing plate at each end of the bundle.
• Insert a soft wood filler board between the bearing plate and tubesheet face to
prevent damage to the tube ends.
• Screw forged steel eyebolts into both bearing plates for pulling and lifting.
• As an alternate to the rods, thread a steel cable through one tube and return
through another tube.
• A hardwood spreader block must be inserted between the cable and each
tubesheet to prevent damage to the tube ends.
7. If the heat exchanger has been in service for a considerable length of time without
being removed, it may be necessary to use a jack on the floating tubesheet to break
the bundle free.
• Use a good-sized steel bearing plate with a filler board between the tubesheet
face and bearing plate to protect the tube ends.
8. Lift tube bundles horizontally by means of a cradle formed by bending a light-gauge
plate or plates into a U-shape. Make attachments in the legs of the U for lifting.
9. Do not drag bundles, since baffles or support plates may become easily bent. Avoid
any damage to baffles so that the heat exchanger will function properly.
10. Some suggested methods of cleaning either the shell side or tube side are listed
below:
• Circulating hot wash oil or light distillate through tube side or shell side will
usually effectively remove sludge or similar soft deposits.
• Soft salt deposits may be washed out by circulating hot fresh water.
• Some commercial cleaning compounds such as ”Oakite” or ”Dowell” may be
effective in removing more stubborn deposits. Use in accordance with the
manufacturer’s instructions.
11. Some tubes have inserts or longitudinal fins and can be damaged by cleaning when
mechanical means are employed. Clean these types of tubes chemically or consult
the nearest manufacturer representative for the recommended method of cleaning.

• If the scale is hard and the above methods are not eﬀective, use a mechanical
means. Neither the inside nor the outside of the tube should be hammered
with a metallic tool. If it is necessary to use scrapers, they should not be sharp
enough to cut the metal of the tubes. Take extra care when employing scrapers
to prevent tube damage.
Do not attempt to clean tubes by blowing steam through individual tubes. This
overheats the individual tube and results in severe expansion strains and leaking
tube-to-tubesheet joints.
12. Table 2 shows safe loads for steel rods and eyebolts.

91
8 Heat exchanger tube side mainenance (Repair vs
replacement
(This subject of chapter is collected from: Bruce W Schafer Framatome ANP, Inc. 155
Mill Ridge Road Lynchburg, VA 24502 (434) 832-3360 bschafer@framatech.com)
Abstract The traditional method of repairing degraded tubes in shell-and-tube heat
exchangers is to remove the effected tubes from service by plugging. Since heat exchangers
are designed with excess heat transfer capability, approximately 10% of tubes can be
plugged before performance is affected. When the number of plugged tubes becomes
excessive, heat exchanger efficiency is lost, resulting in reduced power output, high system
pressure drop, further heat exchanger damage, or abnormal loads placed on other plant
heat exchangers.
As an option to component retubing or replacement, repair methods, including tube sleev-
ing and tube expansion, have proven to be an effective method to repair defective tubes
and keep the existing heat exchanger in service. For the sleeving process, a new tube
section is installed inside the existing tube to bridge across the degraded area. Tube
expansion is used to close off a gap between the tube and the tubesheet or end plate (to
eliminate a leak path) or between the tube and tube support (to minimize vibration).
While not all heat exchangers can be returned to their original design condition by per-
forming tube repairs, in some instances it may be possible to get many more years of
useful life out of a heat exchanger at a fraction the cost of replacement.
This paper presents options which the Plant Maintenance Engineer should consider in
making the repair versus replacement decision. This includes the repair options (sleeving
and tube expansion), other conditions within the heat exchanger, and the effect of tube
repair on heat exchanger performance.
8.1 Introduction
Traditionally, when maintenance is performed on shell-and-tube heat exchangers, the
only options considered when tube defects are found are to plug tubes and, when the
number of plugs became too great, replace the heat exchanger. The decision to replace
the heat exchanger was based on a number of factors. These included: the number of
tubes plugged, the number of forced outages due to tube damage (and the cost associated
with replacing lost power and repairing the damaged tubes), the impact that the plugged
heat exchanger is having on the plant (due to lost flow or heat transfer surface area),
the rate at which tube plugging is occurring, the availability of funds to replace the heat
exchanger, and the expected life of the unit (how much longer will the unit operate before
retirement).
From a sampling of industry data, tube failures have been shown to cause between 31%
to 87% (depending on the data source) of the events related to feedwater heaters (1).
Since so many of the failures were related to the tubing, the replacement of an entire heat
exchanger due to damage in one area is an expensive as well as a schedule and manpower
intensive option.
The typical means for major heat exchanger repair included complete replacement, re-
bundling, and retubing, as described below.
• For the replacement option, the entire heat exchanger shell and tube bundle are
replaced with a new unit.
• For rebundling, the shell is temporarily removed from the heat exchanger and the
old tube bundle, including, at a minimum, tubes, tube supports, and tubesheet, are
removed. A new tube bundle is inserted and the shell is welded back in place.
• For retubing, either the shell (u-tube design) or tube side access cover (straight

92 8 Heat exchanger tube side mainenance (Repair vs replacement
tubes) is removed from the heat exchanger and the old tubes are removed from the
bundle. New tubes are then inserted and re-attached to the tubesheet (typically by
either mechanical expansion, welding, or both). In many instances, the existing shell
side hardware is used as-is, although some modifications may be made. Retubing
is typically performed on straight tube heat exchangers, such as condensers and
coolers.
Since the 1970’s, tube sleeving has been used to allow damaged tubes to remain in ser-
vice. The sleeves are installed by various means (roll, explosive, or hydraulic expansion,
explosively welded, or press-fit or epoxied in place) over the defective area of the tube.
Through the use of sleeving, which is a low-cost option to retubing, rebundling, or re-
placement, the useful life of a heat exchanger can be economically extended. The decision
to perform sleeving also can be made with short notice as opposed to replacement (2-6
weeks compared with 18 months), possibly allowing repairs to be performed the same
outage that the damage is noted. Tube expansion also can be performed to minimize or
eliminate leakage within heat exchangers. In the tubesheet, tubes can be re-expanded to
strengthen the original tube-to-tubesheet joint, reducing or eliminating leakage and pro-
longing the life of the heat exchanger. Expansions also can be made deep within the tube
to expand the tube into tube support plates and end plates. These expansion can reduce
tube-to-plate clearance for vibration control or, at end plates, to minimize steam flow
from the high to low pressure side of the plate.Since the 1970’s, tube sleeving has been
used to allow damaged tubes to remain in service. The sleeves are installed by various
means (roll, explosive, or hydraulic expansion, explosively welded, or press-fit or epoxied
in place) over the defective area of the tube. Through the use of sleeving, which is a low-
cost option to retubing, rebundling, or replacement, the useful life of a heat exchanger
can be economically extended. The decision to perform sleeving also can be made with
short notice as opposed to replacement (2-6 weeks compared with 18 months), possibly
allowing repairs to be performed the same outage that the damage is noted.
Tube expansion also can be performed to minimize or eliminate leakage within heat ex-
changers. In the tubesheet, tubes can be re-expanded to strengthen the original tube-
to-tubesheet joint, reducing or eliminating leakage and prolonging the life of the heat
exchanger. Expansions also can be made deep within the tube to expand the tube into
tube support plates and end plates. These expansion can reduce tube-to-plate clearance
for vibration control or, at end plates, to minimize steam flow from the high to low
pressure side of the plate.
8.2 Repair vs. Replace - Factors To Consider
There are numerous factors to consider when deciding whether to repair the tubes in a
heat exchanger or to perform a larger repair scope and rebundle or replace the component.
The following factors should be considered when making the repair vs. replace decision.
1. The budget available for repair or replacement needs to be determined. Typically,
the cost of performing a substantial heat exchanger repair (consisting of plug re-
moval, tube inspection, tube expansion, and sleeving) is less than 10% of the cost of
replacing the unit. Because of the lower cost, the payback time on the repair option
is much shorter than for replacement.
If the heat exchanger is critical to plant operation (either from a safety, efficiency,
or power production standpoint) or is resulting in costly forced outages, it may be
possible to justify a 3 repair to the unit in the near-term and a scheduled replacement
when a longer outage can be planned. If there are a large number of tube plugs
to remove, or if they are difficult to remove (explosive or welded), then the cost to
repair the heat exchanger will increase, and the scheduled time needed on-site may
not fit within the outage window. If it appears that tube repair may be possible,
it may be worthwhile to plug tubes, using removable plugs, until a certain quantity

8.3 Heat Exchanger maintenance Options 93
of tubes are removed from service. At that point the plugs would be removed and
sleeves installed, thereby minimizing the overall maintenance cost.
2. The location and quantity of the tube defects need to be examined to decide if
tube repair is an option. Tube repair may be appropriate if the damage is limited
to a certain area of the tube, which would allow the use of a short repair sleeve.
If the damage is over a significant portion of the tube, it is possible to install a
longer sleeve (up to the full length of the tube) to ensure that all tube defects are
repaired. However, if the u-bend region of the tube is damaged then tube repair is
not possible. Also, it would not be possible to install a sleeve if a large portion of
the tube had damage but there was inadequate clearance for a long sleeve at the
tube end.
3. One of the more important items to consider when deciding whether a heat ex-
changer can be repaired is the condition of the remainder of the heat exchanger.
The condition of the shell side components, such as the impingement plates, tube
supports, end plates, and other structural members, should be in good shape if a
long term repair is being planned. An evaluation also should be made of the shell
thickness in areas that are prone to shell erosion/corrosion. If the tube repair is only
a short-term fix, to allow component operation until a replacement heat exchanger
can be installed, the condition of the shell side is not as critical.
4. The life expectancy of the power plant needs to be factored into the decision to
repair or replace a heat exchanger. If the only problem with the heat exchanger is
in one section of the tube, and the expected run time on the unit is relatively short,
it would be advantageous to repair rather than replace the heat exchanger since it
will be very difficult to pay back the cost for replacement over the remaining plant
life.
5. The outage time required to repair a heat exchanger, even when tube and shell side
inspections are performed, is typically much less than for replacement. In addition,
very few, if any, plant modifications need to be made to make the repairs. This
allows other work to be performed in the vicinity of the heat exchanger. Along
with the shorter outage duration, the site support required for repair is much less.
Usually, there are no shell or head modifications required since all work can usually
be performed through the manways and pass partition plates. Less repair equipment
is required, resulting in less space being needed in the area of the heat exchanger
for setup and storage. In addition, the time required to prepare for tube repair is
much less than for replacement (2- 6 weeks compared with 18 months), allowing a
decision on repair to be made just before, or even during, an outage.
6. At nuclear plants, the added cost for the disposal of radioactively contaminated
heat exchangers must be taken into account. Before disposal, there is the cost of
surveying the heat exchangers for release and, if contamination is found, they must
either be decontaminated or disposed of as radioactive waste. Tube repairs can
eliminate these costs.
7. If the heat exchanger is being replaced to eliminate detrimental materials in the
cooling system (i.e. copper in the condensate/feedwater system) then tube sleeving
will not be beneficial. The only solution would be to retube/rebundle/replace to
change out the tube material.
8.3 Heat Exchanger maintenance Options
There have always been options available to either repair or replace heat exchanger tubes
in the event that tube leakage or degradation is present. The repair options include:
1. Plug the tube

2. Sleeving
3. Tube expansion
The replacement options include
1. Retubing
2. Rebundling
3. Replace with new unit
8.4 Repair option
8.4.1 Plug
The initial option, after the problem tubes have been located (either through non-destructive
examinations, such as eddy current testing, visual inspections, or leak tests) is to plug the
tube. Depending on the type of service and operating pressures of the heat exchanger,
various types of plugs are employed. These include
1. tapered fiber and metal pin plugs,
2. rubber compression plugs,
3. two piece ring and pin plugs,
4. two piece serrated ring and pin plugs (installed with a hydraulic cylinder),
5. welded plugs, and explosively welded plugs.
In addition to the tube end plug, there also may be a stabilizer rod or cable that is inserted
into the tube to minimize future tube vibration damage.
At the beginning of the life of a heat exchanger, inserting a few plugs into damaged tubes
has little effect on the performance of the heat exchanger. However, if heat exchanger
problems continue, and the number of plugs increases significantly, it is possible that
the heat exchanger will eventually reach a point that it will not handle the full load
that is placed on it. This is due to a combination of loss of heat transfer area and the
increased pressure drop. In addition, as the number of plugged tubes increases, abnormal
temperature conditions (either hot or cold spots) may be set up in the heat exchanger.
These conditions can result in an acceleration of tube damage, creating a faster demise
of the heat exchanger.
Once the number of plugs reaches a unacceptable level, the heat exchanger will need to be
repaired, replaced, or bypassed. However, bypassing the unit is usually not recommended,
at least for a long time period, since it will result in a loss of efficiency and heat transfer
area. Also, the heat load from the bypassed heat exchanger will be transferred to another
heat exchanger in the string, resulting in greater than normal operating flow rates and
higher degradation in that heater.
The following sections show the options that can be used to replace or repair the entire
heat exchanger or just the tubes.

8.4 Repair option 95
8.4.2 Sleeving
An alternate approach to retubing, rebundling, or replacement of a heat exchanger is to
install sleeves over the defective portions of the tubes. The sleeve consists of a smaller
diameter piece of tubing that is inserted into the parent tube and positioned over the
tube defects. After insertion, each end of the sleeve is expanded into the parent tube
material. These expansions serve the dual function of structurally anchoring the sleeve
into the tube and providing a leak limiting path, allowing the sleeve to become the new
pressure boundary for the tube. This means that a sleeved tube can have a 100% through-
wall indication and still remain in-service, since the sleeve is now the new structural and
pressure boundary. The installation of the sleeve into the tube will allow the majority of
the tube’s heat transfer area and ﬂow to be maintained.
If heat exchanger repair by sleeving is a possibility then a strategy needs to be used to
prepare for future repair. It may be cost eﬀective to plug a quantity of tubes, per the non-
destructive examination results, each outage using a removable plug. When the quantity
of plugged tubes reaches a certain level the plugs can be removed and sleeves installed.
Using this approach will minimize the cost and time during each inspection outage while
allowing the maximum tube repair later in the heat exchanger’s life.
There are three types of sleeves that are installed into heat exchanger tubes. These are
1. full length,
2. partial length structural, and
3. partial length barrier sleeves.
The three types are discussed below. Figure Figure 8.1 shows the sleeve layout.
Figure 8.1. Heat Exchanger Sleeve Designs

Full Length Sleeve
These sleeves are installed from one end of the tube to the other in straight tubed heat
exchangers. After insertion, the full length of the sleeve is expanded into the parent
tube. This step serves the dual purpose of maintaining heat transfer as high as possible
(typically 75%-90%) while minimizing flow pressure drop through the tube. After the full
length expansion step, shown in Figure 8.2, the sleeve ends are trimmed flush with the
existing tube ends and the sleeve is roll expanded into the tubesheet.
The full length sleeve is typically used in a condenser or cooling water heat exchanger when
the tubes have multiple defects along their length. Full length sleeving is an attractive
option when a relatively small percentage of the tubes require repair. Through sleeving,
the majority of the tube heat transfer area can be left in service, resulting in a heat
exchanger that is close to its as designed condition.
Full length sleeving is comparable in many ways to retubing in the methods employed to
install the sleeves. However, since removal of the existing tube is not required, and the
typical number of tubes that will be full length sleeved are below the number that would
be retubed, the cost for material and manhours are much less than for retubing, making
sleeving a cost-effective option to return and keep tubes in service.
Figure 8.2. Full Length Sleeve Expansion
Partial Length Structural Sleeve
This type of sleeve is used to repair shorter defects in the tube. The sleeve can be
installed anywhere along the straight length of the tube. Various methods are used to
expand the sleeve in place. These include roll expansion (both in the tubesheet and in
the freespan portion of the tube), hydraulic expansion in the freespan portion of the tube,
and full length expansion. These expansion types are discussed below. The installation
of a hydraulically expanded sleeve is shown in Figure 8.3.
• If one end of the sleeve is in the tubesheet, a torque-controlled roll expansion will be
made. This expansion is similar to the original tube-to-tubesheet roll. Freespan roll
expansions are made to either a torque controlled setting or to a diameter controlled
hardstop setting. Usually, freespan roll expansions are only used when the sleeve
length is relatively short, since it can be difficult to insert a roll expander deep into
the tube. Both the tubesheet and freespan roll expansion parameters are set so that
they can provide both the structural and leakage requirements for the sleeve.

• For sleeves installed deep within the tube, a hydraulic expansion device is used to
connect the sleeve to the tube. The expander consists of multiple plastic bladders
that are filled with high pressure water. As the water pressure increases, the bladders
expanded against the inside of the sleeve, pushing the sleeve into the tube. The
expansion process, which is computer controlled, continues until either a preset
volume of water or a preset pressure is reached. At this point the sleeve is properly
expanded and the bladders are depressurized. Hydraulic expansions can be made
anywhere along the tube length since the expander is connected to flexible high
pressure tubing and is not restricted by tube end access. The expansion parameters
are qualified to meet the proper structural and leakage requirements for the sleeve.
• Full length expansions are not usually used for structural or leak limiting purposes
but instead are used to improve heat transfer and flow through the sleeve and to
close the annulus between the sleeve and tube. The full length expansion is made
by placing a tool, with seals on each end, into the sleeve. The inside of the sleeve
is filled and then pressurized with water to a preset pressure setting, expanding the
sleeve into tight contact with the tube. After the full length expansion is made,
the ends of the sleeve are typically either roll or hydraulically expanded to form the
structural and leak limiting sleeve-to-tube joint.
Many times, the partial length structural sleeves are used to repair indications at one
particular area of the tube, such as wear damage at tube support locations, cracking
in roll transitions, or pitting indications at one discreet location along the tube length.
Longer versions of these sleeves also have been used to repair an entire damaged section
of a heat exchanger, such as a desuperheater or drain cooler section of a feedwater heater.
Because of the wide variety of uses, the sleeve length can range from as short as 1 foot to
over 12 feet in length.
Qualification testing is performed on the structural sleeves to ensure that they can with-
stand the design temperature and pressure conditions imposed on them. The test results
must show that the sleeve will be the new pressure boundary even with a 100% through-
wall indication in the parent tube. Sleeves of this type, using mechanical expansions (roll
and hydraulic), have reliably been in-service for more than 15 years.
Partial Length Barrier Sleeve
These sleeves, also known as shields, are used at the ends of the tubes to act as a barrier
to tube end erosion. These sleeves are usually very thing, are not designed to act as a
pressure boundary or structural repair, and are installed in areas of high turbulence. The
materials for these sleeves are compatible with the existing tube material and may include
plastic inserts. The sleeves are either roll or hydraulic expanded or pressed or epoxied
in place. If tube end erosion is occurring, or is expected to occur, the use of these tube
end sleeves will protect and prolong the life of the parent tube, although over time tube
erosion may begin to occur at the end of the sleeve. Many heat exchanger tube ends have
been protected with shields, significantly prolonging the life of the tubes.
Items to Consider for Tube Sleeving
Prior to choosing to perform tube sleeving, the following factors should be considered.
• The length, location, and quantity of tube defects that would require sleeving need
to be determined. If the defects are in one or a few short areas then either a single or
a couple of partial length sleeves could be used. However, if the defects are spaced
throughout the length of the tube, then the only option would be a full length sleeve.
The parent tube in the area where the sleeve will be expanded is to be defect free.

Figure 8.3. Partial Length - Hydraulically Expanded Structural Sleeve Installation
This will insure the highest sleeve-to-tube joint integrity. Also, the tube support
designations must be clearly identified to insure that the sleeve is installed at the
correct location along the tube length. This is especially true in areas where there
may be skipped baffles and the tube only touches every other support plate.
• The condition of the remainder of the tube away from the sleevable defects needs
to be known. If there are u-bend defects that may require plugging then the tube
should not be sleeved. Sleeving is an option if the remainder of the tube is in good
shape.
• The space available at the tube end to insert a sleeve and its installation tooling
needs to be known, as shown in Figure 8.4. If a short, partial length sleeve is being
used, the amount of space required is not as critical, although there can still be
access issues around the tubesheet periphery for hemi-head channel covers and at
pass partition plates. However, if a full length sleeve is required, there will need to
be a significant amount of clearance from the tubesheet face.
• Inspection records need to be reviewed to determine if there are any tube inside
diameter (ID) restrictions that would block the sleeve from being inserted to the
target location. The size of the eddy current probe used for the inspection, plus any
other hardware that has been inserted into the tube, can be used to help determine
the tube ID access issues.
• The post-sleeving tube inspection requirements need to be considered. Typically,
the ability to inspect the tube beyond a sleeve is not a significant issue. While
the presence of the sleeve reduces the inside diameter of the tube, which will result
in the need for a smaller inspection probe, the probe will remain large enough
to detect pluggable tube indications (usually greater than 40%), however small

indications may go undetected. As part of the post-sleeve inspection, the sleeve and
its attachment to the tube should be examined. There is no need to inspect the
section of the parent tube between the sleeve expansions since this is no longer part
of the pressure boundary.
• If tube cleaning is to be performed in the heat exchanger, then the type of sleeve to
be installed needs to be evaluated. If on-line cleaning is performed, the sleeve size
cannot restrict the passage of the balls or brushes. For off-line cleaning, the projec-
tiles need to pass through the sleeve without becoming stuck. Many sleeves that are
installed in tubes that require cleaning are full length expanded to ensure the best
results for the cleaning equipment. If it appears that tube sleeving is possible, then
information will be needed to ensure that the heat exchanger is properly repaired.
The following information is used when planning for sleeving.
• Tube sleeving will need to be coordinated with eddy current inspection and plug
removal.
• If it is expected that sleeving may be performed, then it is important that the proper
sleeve material be purchased in advance of the job.
• The sleeve material needs to be compatible with the heat exchanger parent tubing
and with the water chemistry within the heat exchanger. The galvanic corrosion
potential between the sleeve and tube needs to be determined. Also, effects of crevice
corrosion between the sleeve and tube, in the heat exchanger water chemistry, need
to be considered to determine if sleeving is a viable repair option.
• The sleeve dimensions need to fit the heat exchanger operating and design condi-
tions plus any restrictions within the tube ID. The sleeve outside diameter (OD) is
to be designed to fit into the tube but must be long enough to limit the amount of
sleeve expansion. The sleeve wall thickness needs to be sized for the heat exchanger
operating parameters, including any ASME Code minimum wall thickness calcula-
tions, if needed. The sleeve length must be long enough to span the expected tube
defects but needs to be sized to fit any tube end clearance restrictions.
• Before installing sleeves into heat exchanger tubes, testing needs to be performed to
set the installation parameters. Depending on the type of sleeve being used, these
tests may include setting the rolling torque, hydraulic expansion constants, and full
length expansion pressure. In addition, depending on the application for the sleeve,
there may be a need to do qualification testing, which would consist of hydrostatic
leak and pressure tests and temperature and pressure cycling. These tests would
verify that the expansion parameters were set correctly for the sleeve application.
• If a large quantity of sleeves are being installed, it may be necessary to calculate
the heat transfer and flow loss due to sleeving. These calculations will give a sleeve-
to-plug ratio that can be used to determine the expected improvement in heat
exchanger performance after sleeving is complete (and tubes have been returned to
service, if applicable).
• The sleeve may need to be full-length expanded based on the heat exchanger oper-
ating environment. However, the production rates for sleeve installation are lower
when full length expansions are performed. While full length expansion is typi-
cally not needed in many applications, such as most feedwater heaters, it should be
considered for the following.
– if tube ID cleaning needs to routinely be performed
– if a long sleeve is being inserted that would severely restrict the tube’s heat
transfer or flow

– if the tube-to-sleeve crevice needs to be eliminated in a hostile water chemistry
environment
– if there are large eddy current probe fill factor restrictions
Figure 8.4. Required Clearance for Sleeve Installation
8.4.3 Tube Expansion
In addition to sleeving, it is possible to expand the tube to improve the heat exchanger
performance. These tube repairs can minimize further tube damage and maximize the
useful life of the heat exchanger. Two methods of tube expansion can be performed. One
is to expand deep within the tube to close off a leak path between the tube and the end
plate. The other is to re-expand the tube into the tubesheet to minimize tube-to-shell
side leakage.
Tube-to-End Plate Expansion
In some heat exchangers, typically feedwater heaters, there are internal plates which
separate one zone of the heat exchanger from another (usually condensing [steam] from
drain cooler [liquid]). Due to the pressure differential across the plate, and the different
temperatures and phases between the two sections, it is important that leakage not occur
through the plate. However, in some feedwater heaters, the plate design is too thin,
resulting in leakage of steam from the condensing to the drain cooler zones, as shown in
Figure 8.5. When this occurs there is erosion of the end plate and tube vibration due to
the high steam velocities and the steam condensing to liquid in the drain cooler region.
The vibration causes wear at the tube supports which can lead to tube failure. The
leakage of steam also increases the drain cooler temperature, resulting in a less efficient
heat exchanger. Expanding the tube can reduce the gap between the tube and the end
plate. The expansion can be performed using either a roll or hydraulic expander. Once the
expander is in position the tube is expanded until it contacts the end plate. An accurate
expansion, which does not over-expand the tube into the plate (the tube needs to be able
to slide in the plate after expansion so that it does not buckle during heatup/cooldown),

Figure 8.5. Required Clearance for Sleeve Installation
needs to be performed. This can be achieved by using a computer controlled hydraulic
expansion that automatically shuts off the pressurization system when it detects that the
tube has contacted the plate.
After the tubes are expanded into the end plate, the steam flow is minimized or elimi-
nated, reducing the drain cooler temperatures and increases plant efficiency. Further tube
damage, in the form of tube wear and adjacent tubes impacting on one another, will be
reduced to nearly zero and the vibration operating stresses will be reduced significantly.
The life of the heat exchanger will be increased at a minimal cost as compared with
replacement.
Tube-to-Tubesheet Expansion
In some heat exchanger designs, with a certain combination of materials, leaks develop
between the tube and tubesheet. In many low pressure units, the tube is only expanded
into the tubesheet, with no subsequent weld. Many of the leaks that occur in these units
are the result of a fabrication error and can be corrected by re-expanding the joint to
the correct expansion size. However, leakage occasionally occurs in high pressure heat
exchangers, typically feedwater heaters, even when the tubes have been welded to the
tubesheet. The two prime causes of this leakage are in areas where the original tube-to-
tubesheet weld has either cracked or eroded due to flow (in the case of soft materials, such
as carbon steel) or where there is a crack in a tube-totubesheet expansion transition.
• For the first case it may be possible to re-expand the tube using a qualified roll
expansion process. The expansion would increase the contact pressure between the
tube and tubesheet, increasing the resistance to flow and decreasing or eliminating

leakage. This process could be performed on existing leaking tubes or preventatively
on all tubes in the tubesheet.
• If cracking is occurring at the original tube expansion transition it may be possible
to re-expand the tube deeper in the tubesheet (unless the cracking is occurring
very close to the shell side of the tubesheet). The tube would be expanded using
a qualified roll expansion process, to place the tube into tight contact with the
tubesheet. This expansion would increase the contact pressure between the tube and
tubesheet, increasing the resistance to flow and decreasing or eliminating leakage.
This process could be performed either on existing leaking tubes or preventatively
on all tubes in the tubesheet.
Re-expanding tubes that either may be leaking or that could develop leaks in the future
could significantly extend the life of an otherwise good heat exchanger. By re-expanding
the tubes, forced outages can be avoided and damage from the high pressure water spray-
ing on adjacent tubes and on the shell will be eliminated. The cost to perform tube
re-expansions will be minimal when compared with the cost of replacement heat exchang-
ers and the cost of forced outages.
Items to Consider for Tube Expansion Repair
The following factors should be considered to determine if tube expansion is possible.
• The portion of the tube to be expanded needs to be determined.
– If leakage is occurring through the end plate, the expander will need to be long
enough to reach the end plate location. The tube should be expanded using a
process, such as hydraulic expansion, that will not lock the tube into the end
plate. This expansion will not only reduce leakage through the plate but also
will minimize future tube vibration due to the tight fit between the tube and
plate.
– If leakage is occurring within the tubesheet, due to either weld or tube cracking,
a re-expansion process may be used. This process, typically a roll expansion,
will reexpand the tube into the tubesheet to limit or eliminate leakage from
the tube to the shell side of the heat exchanger.
• The condition of the remainder of the tube needs to be known. If there are cracks
along the entire tube length then re-expanding the tube alone will not result in an
improvement in heat exchanger performance.
• The space available at the tube end to insert the expansion tooling needs to be
known. Usually either a roll or hydraulic expander will be used for this process.
Unless a roll expansion is being performed at the end plate, the usual repair tooling
is relatively short, although there can still be access issues around the tubesheet
periphery for hemi-head channel covers and at pass partition plates.
• For tube end plate expansions, the eddy current inspection records need to be
reviewed to determine if there are any tube inside diameter restrictions that would
block the expander from being inserted to the end plate location. The size of the
eddy current probe used for the inspection, plus any other hardware that has been
inserted into the tube, can be used to help determine the tube ID access issues.
The potential for any tube end restrictions, that might limit tooling insertion into
the tube, also need to be known so that tooling can be prepared to eliminate the
restriction.

8.5 Replacement option 103
If it appears that tube expansion is possible, then information will be needed to ensure
that the heat exchanger is properly repaired. The following information is used when
planning for tube expansion.
• Tube expansion will need to be coordinated with eddy current inspection and plug
removal.
• The tube expander design (diameter and length) needs to be based on the require-
ments for the expansion. Before performing tube expansions into heat exchanger
tubes, testing needs to be performed to set the tooling operating parameters. De-
pending on the type of expansion, these tests may include setting the rolling torque
for tubesheet re-expansions or setting the hydraulic expansion constants for end
plate expansions. In addition, for the tube-intotubesheet re-expansion process, qual-
iﬁcation testing should be performed. This would consist of hydrostatic leak and
pressure tests and temperature and pressure cycling. These tests would verify that
the expansion parameters were set correctly for the tube reexpansions. exchanger.
8.5 Replacement option
8.5.1 Retubing
The tubes can be replaced, if the unit has:
• straight tubes,
• good access, and
• the remaining components (shell, tube supports, internal structural pieces) of the
heat exchanger are in good shape.
The old tubes are removed from the unit and new ones, typically manufactured from
an improved material, are inserted, and then expanded, into place. Insertion of the new
tubes is shown in Figure 8.6. In addition to performing retubing to replace damaged
tubes, retubing has been performed to eliminate detrimental materials (such as copper
from condenser tubes) to minimize damage to other equipment within the plant (nuclear
steam generators or fossil boilers).
Figure 8.6. Condenser Retubing

8.5.2 Rebundling
Some heat exchangers are designed to be rebundled rather than replaced. For these units
the entire tube bundle, including tubes, tubesheet, and tube supports are replaced, as
shown in Figure 8.7. The original shell and any other internal structural pieces would
be reused (although any necessary internal repairs could be made when the shell was
removed). The new tube bundle can be manufactured to ensure that original design
problems with the existing unit are corrected. However, the same basic design must
be maintained since the new bundle must ﬁt within the existing heat exchanger shell.
Rebundling costs about 15-25% more than retubing.
Figure 8.7. Heat Exchanger Rebundling
8.5.3 Complete replacement (New unit)
A third and typically widely used option is to replace the entire heat exchanger, as shown
in Fig.8.8 . Full replacement allows alternate tube materials, changes in heat transfer area,
and structural changes to be employed, including added clearances in areas where erosion
or other problems may be occurring, to ensure that the current heat exchanger problems
do not re-occur in the future. However, the cost associated with a full replacement is the
greatest of the three options, about 5% more than for rebundling . In addition, there
are no guarantees that the new heat exchanger design will not have new, unanticipated
problems.
Figure 8.8. Heat Exchanger Replacement

8.6 Conclusions 105
8.6 Conclusions
The costs associated with heat exchanger replacement can be significant. These costs
include the new heat exchanger or tube bundle, the manpower required to remove the
old and install the new heat exchanger components, plant modifications to allow for the
removal of the heat exchanger, and the amount of outage time associated with replace-
ment. In addition, the replacement of a heat exchanger can adversely affect other work
going on in the their vicinity. Because of the cost and time involved, and if the damage
is confined to only the tubing (which is typically the case), repair of the heat exchanger,
through either sleeving or tube expansion, should be considered. If the tube damage is
confined to one general area, there is a good possibility that the expense of a replacement
can be avoided. In addition, the time required to prepare for tube repair is much less
than for replacement (2-6 weeks compared with 18 months), allowing a decision on repair
to be made just before, or even into, an outage.
By removing plugs and installing sleeves, it is possible to return lost heat transfer area to
service. Tubes that would be likely to fail in the near term also can be repaired. This will
improve the performance and reliability of the heat exchanger. The cost to perform the
repairs is also much less than for replacement (usually less than 1/10th the cost). Sleeving
has been shown to be a proven tube repair technique, having been performed since the
1970’s. During this time, tube repairs have economically extended the useful life of heat
exchangers worldwide.
As the number of plugged tubes approaches the upper limits or if damage is consistently
occurring in one area of a heat exchanger, tube repair, through both sleeving and tube
expansions, should be considered to minimize future damage and extend the life of the
heat
The following table shows the various heat exchanger repair options and the factors to
be considered when choosing each of the options. Note that the table contains selected
criteria for evaluating component repair versus replacement options. A final decision to
implement a particular option should be made on a case by case basis with proper weight
given to all factors. The information listed in this table is for relative comparison purposes
only.

106 9 Troubleshooting
9 Troubleshooting
9.1 Heat exchangers’ problems
Heat transfer equipment provides the economic and process viability for many plant op-
erations. The basis for successful application of such equipment depends on the designer.
The problem that should be anticipated by the design to avoid high maintenance or
cleaning and costly shut down production include:
1. Fouling
2. Leakage
3. Corrosion
To anticipate maintenance problems the designer should need to be familiar with the
plant location, process flow sheet, plant operation. Some of the questions that must be
considered are:
1. will the heat exchanger need cleaning? how often? what cleaning method will be
used?
2. what penalty will the plant pay for leakages between the tubeside and shell side?
3. what kind of production upsets can occur that could affect the heat exchanger?will
cycling occur?
4. how will heat exchanger be started up and shut down?
5. will the heat exchanger be likely to require repairs? if so, will the repairs present
any special problem?
9.2 Fouling
9.2.1 Costs of fouling
• Increased maintenance costs
• Over-sizing and/or redundant (stand-by)equipment
• Special materials and/or design considerations
• Added cost of cleaning equipment ,chemicals
• Hazardous cleaning solution disposal
• Reduced service life and added energy costs
• Increased costs of environmental regulations
• Loss of plant capacity and/or efficiency Loss of waste heat recovery options

9.2 Fouling 107
9.2.2 Facts about fouling
• 25 YEARS AGO heat exchanger fouling was referred to as ”the major unresolved
problem in heat transfer” ?
• the total cost of fouling - in highly industrialized nations - has been projected at
0.25% of the GNP ?
• the total annual cost of fouling in the U.S. is now estimated at 18 billion ?
• the total annual cost of fouling specifically focused on shell and tube exchangers in
the process industries is now estimated at 6 billion ?
9.2.3 Types of Fouling
• Precipitation / Crystallization - dissolved inorganic salts with inverse solubility char-
acteristics
• Particulate / Sedimentation - suspended solids, insoluble corrosion products, sand,
silt
• Chemical Reaction - common in petroleum refining and polymer production
• Corrosion - material reacts with fluid to form corrosion products, which attach to
the heat transfer surface to form nucleation sites
• Biological - initially micro-fouling, usually followed by macro-fouling
• Solidification - ice formation, paraffin waxes
9.2.4 Fouling Mechanisms
• Initiation - most critical period - when temperature, concentration and velocity
gradients, oxygen depletion zones and crystal nucleation sites are established - a
few minutes to a few weeks
• Migration - most widely studied phenomenon - involving tranport of foulant to
surface and various diffusion transport mechanisms
• Attachment - begins the formation of the deposit
• Transformation or Aging - another critical period when physical or chemical changes
can increase deposit strenght and tenacity Removal or
• Re-entrainment - dependent upon deposit strength - removal of fouling layers by
dissolution, erosion or spalling - or by ”randomly distributed turbulent bursts”
9.2.5 Conditions Influencing Fouling
• Operating Parameters
1. velocity
2. surface temperature
3. bulk fluid temperature
• Heat Exchanger Parameters
1. exchanger configuration
2. surface material

3. surface structure
• Fluid Properties
1. suspended solids
2. dissolved solids
3. dissolved gases
4. trace elements
9.2.6 Fouling control
1. Good design:
(a) Forced circulation heat excahnger. Forced circulation calandria is better than
natural circulation calandria. This is to obtain a velocity of 10-15ft/sec. Al-
though the cost of pumps and power added considerably to the cost of the
equipment. This would be compared to the cost of production losses and cost
for cleaning in order to arrive to at an economical design for a particular process
application.
(b) Good shell side avoids eddies and dead zones where solid can accumulate. Inlet
and outlet connections should be located at the bottom and top of the shell
side and tube side to avoid creating dead zones and unvented areas.
(c) The use of metal that will not foul due to accumulation of corrosion products is
important, especially with cooling waters. Copper, copper alloy and stainless
steels are satisfactory for most cooling waters
2. The fouling fluid should be inside tube. Hence easily removable flat cover plates
would be installed on the channel to facilitate cleaning if frequent physical cleaning
is necessary. Horizontal installation would probably be chosen to avoid the cost of
scaffold usually required for physically cleaning a vertical exchanger
3. Increasing tube velocity to 10-15ft/s lengthen the cleaning intervals
4. Using heat transfer equipment with single flow channel will often reduce fouling due
to sedimentation. For example spiral plate heat exchanger may be selected in place
of a multipass shell and tube heat exchanger unit to avoid settling of suspended
solids in the shell side and at the bottom of the tube side bottom of the tube side
channel.
9.2.7 Fouling cleaning methods
1. Chemical cleaning: Various chemicals (acids, chlorine) have been used to reduce
fouling and restore tube cleanliness. Acid may either be strong (which damage the
equipment) or week (citric, formic, sulfamic) these are less effective. Acid cleaning
is limited to once a year or less. The use of chlorine is being cutback or eliminated
in many regions by government regulations.
2. Manual cleaning. Method include periodic cleaning with rubber plugs, nylon brushes,
metal scrapers or turbining tools. This method is expensive, intermittent (between
cleaning fouling builds up rapidly)
3. Rubber - ball cleaning: Automatic cleaning by means sponge -rubber balls is eco-
nomical in areas where deposition, pollutants, chlorides and other corrodents exists.
These ball distribute themselfs at random through the condenser, passing through
a tube at an average of one every five minutes. slightly larger in diameter than the
tube, they wipe the surface clean of fouling and deposits

9.3 Leakage/Rupture of the Heat Transfer Surface 109
9.3 Leakage/Rupture of the Heat Transfer Surface
Leaks may develop at
1. the tube-to-tubesheet joints of fixed tube sheet exchanger due
(a) to differential thermal expansion between the tube and shell causes overstress-
ing of the rolled joints, or
(b) thermal cycling caused by frequent shutdowns or batch operation of the process
may cause the tubes to loosen in the tube holes.
2. Leaks may occur due to tube failure cause by vibration or differential thermal ex-
pansion or dryout (for boilers and evaporators)
9.3.1 Cost of leakage
1. Large production losses or maintenance cost
2. Contamination of product:The leak/rupture of tubes leads to contamination or over-
pressure of the low-pressure side. Failure to maintain separation between heat trans-
fer and process fluids may lead to violent reaction in the heat transfer equipment
or in the downstream processing equipment.
9.3.2 Cause of differential thermal expansion
1. Unusual situation that lead to unexpected differential thermal expansion, for ex-
ample, the tube side of a fixed-tube sheet condenser may be subjected to steam
temperature, with no coolant in the shell whenever a distillation column is steamed
out in preparation for maintenance. Or an upset in the chemical process may subject
the tubes to high temperatures
2. Start up at high temperature
3. Vibration (if the velocity at the inlet exceeded the critical velocity for two phase
flow)
4. Dryout of the tube cause by insufficient coolant or local overheating
Remedy of thermal expansion
1. Use of U tube or floating head instead of fixed tube sheet
2. Welding the tube to the tube sheet
3. Double tube sheet
4. Use large nozzle or vapor belts to give velocity well below the critical
To make the heat transfer process inherently safer, designers must look at possible in-
teractions between heating/cooling fluids and process fluids. For relatively low-pressure
equipment (<1000 psig), a complete failure of tubes may not be a credible overpressure
scenario if the design pressure of the low-pressure side and associated equipment is greater
than two-thirds of the design pressure of the high- pressure side (API RP 521 1993), or if
the geometry of the tube layout is such that a complete break is not physically possible.
For high-pressure equipment (> 1000 psig), however, a complete failure should be consid-
ered credible, regardless of pressure differential.

9.4 Corrosion
The heat transfer surface reacts chemically with elements of the fluid stream producing
a less conductive, corrosion layer on all or part of the surface.
9.4.1 Corrosion effects
1. Premature metal failures
2. the deposit of corrosion products reduce both heat transfer and flow rate.
9.4.2 Causes of corrosion
High content of total dissolved solids (TDS), the dissimilarity of the metal, dissolved
oxygen, penetrating ions like chlorides and sulphates, the low pH and presence of various
other impurities are the prime cause of corrosion in the heat exchanger.
9.4.3 Type of corrosion
• stress corrosion
• galvanic corrosion
• uniform corrosion
• Pitting
• Crevice Corrosion
9.4.4 Stress corrosion
• Differential expansion between tubes and shell in fixed-tube-sheet exchangers can
develop stresses, which lead to stress corrosion.
• Overthinning: Expanding the tube into the tube sheet reduces the tube wall thick-
ness and work-hardens the metal.
• The induced stresses can lead to stress corrosion.
Controlling Stress Corrosion Cracking
• Proper selection of the appropriate material.
• Remove the chemical species that promotes cracking.
• Change the manufacturing process or design to reduce the tensile stresses.
9.4.5 Galvanic corrosion
Galvanic corrosion is frequently referred to as dissimilar metal corrosion. Galvanic corro-
sion can occur when two dissimilar materials are coupled in a corrosive electrolyte.

9.4 Corrosion 111
9.4.6 Pitting
Pitting is a localized form of corrosive attack. Pitting corrosion is typified by the formation
of holes or pits on the tube surface.
Causes:
• dissolved oxygen content
• eposition of corrosion products
Methods for reducing the effects of pitting corrosion: Reduce the aggressiveness of the
environment (pH, O2) Use more pitting resistant materials Improve the design of the
system
9.4.7 Uniform or rust corrosion
Some common methods used to prevent or reduce general corrosion are listed below:
• Coatings
• Inhibitors
• Cathodic protection
• Proper materials selection
9.4.8 Crevice corrosion
Crevice corrosion is a localized form of corrosive attack. Crevice corrosion occurs at
narrow openings or spaces between two metal surfaces or between metals and nonmetal
surfaces.Some examples of crevices are listed below:
• Flanges
• Deposits
• Washers
• Rolled tube ends
• Threaded joints
• O-rings
• Gaskets
• Lap joints
• Sediment
Some methods for reducing the effects of crevice corrosion :
• Eliminate the crevice from the design. For example close fit. A 3-mm- long gap is
thus created between the tube and the tube hole at this tube-sheet face. The tube
is allowed to protrude 3 mm of the tube sheet.
• Select materials more resistant to crevice corrosion
• Reduce the aggressiveness of the environment

9.4.9 Materials of Construction
The various parts of the heat exchanger (tube, shell, tube sheet, baffles, front head, rear
head, nozzles,...) may be manufactured from same metal or dissimilar metals. Individual
components may be fabricated from single metal or bimetallic.
For the selection of material of construction, the corrosion chart must be consulted (Ap-
pendix C of Coulson and Richardson [29]). The chart gives metal (alloy) vs chemical at
various temperatures. Note:Before using the corrosion chart the notation given should
read thoroughly.
9.4.10 Fabrication
Expanding the tube into the tube sheet reduces the tube wall thickness and work-hardens
the metal. The induced stresses can lead to stress corrosion. Differential expansion
between tubes and shell in fixed-tube-sheet exchangers can develop stresses, which lead
to stress corrosion.
When austenitic stainless-steel tubes are used for corrosion resistance, a close fit between
the tube and the tube hole is recommended in order to minimize work hardening and the
resulting loss of corrosion resistance. In order to facilitate removal and replacement of
tubes it is customary to roller-expand the tubes to within 3 mm of the shellside face of
the tube sheet. A 3-mm- long gap is thus created between the tube and the tube hole
at this tube-sheet face. In some services this gap has been found to be a focal point for
corrosion.
It is standard practice to provide a chamfer at the inside edges of tube holes in tube sheets
to prevent cutting of the tubes and to remove burrs produced by drilling or reaming the
tube sheet. In the lower tube sheet of vertical units this chamfer serves as a pocket
to collect material, dirt, etc., and to serve as a corrosion center. Adequate venting of
exchangers is required both for proper operation and to reduce corrosion.
Improper venting of the water side of exchangers can cause alternate wetting and drying
and accompanying chloride concentration, which is particularly destructive to the series
300 stainless steels.
Certain corrosive conditions require that special consideration be given to complete drainage
when the unit is taken out of service.
Particular consideration is required for the upper surfaces of tube sheets in vertical heat
exchangers, for sagging tubes, and for shell-side baffles in horizontal units.
9.5 Troubleshooting
This chapter presents potential failure mechanisms for heat transfer equipment and sug-
gests design alternatives for reducing the risks associated with such failures. The types
of heat exchangers covered in this chapter include:
• Shell and tube exchangers
• Air cooled exchangers
• Direct contact exchangers
• Others types including helical, spiral, plate and frame, and carbon block exchangers
This chapter presents only those failure modes that are unique to heat transfer equipment.
Some of the generic failure scenarios pertaining to vessels may also be applicable to heat
transfer equipment. Unless specifically noted, the failure scenarios apply to more than
one class of heat transfer equipment.

9.6 Past failure incidents 113
9.6 Past failure incidents
This section provides several case histories of incidents involving failure of heat transfer
equipment to reinforce the need for the safe design practices presented in this chapter.
9.6.1 Ethylene Oxide Redistillation Column Explosion:
In March 1991, an Ethylene Oxide (EO) redistillation column exploded at a Seadrift,
Texas chemical facility. The explosion was caused by energetic decomposition of essen-
tially pure EO vapor and liquid mist inside the column.
A set of extraordinary circumstances was found to have coincided, resulting in the catalytic
initiation of decomposition in a localized region of a reboiler tube. Extensive investigation
by reference [158] showed that:
1. A low liquid level in the column, plus a coinciding temporary condensate backup
and accumulation of inert gas in the reboiler shell, significantly diminished the EO
liquid fraction leaving the reboiler. Nevertheless, sufficient heat transfer capacity re-
mained to satisfy the vaporization rate required by the column controls, so operation
appeared normal.
2. A localized imbalance resulted in some reboiler tubes losing thermosyphon action,
so that the existing EO was essentially all vapor. Due to ongoing reaction with
traces of water, high boiling glycols accumulated in the stalled tubes, increasing
the boiling point while reducing the heat flux and resulting mass flow rate. This
self-reinforcing process continued leading to minimal EO vapor velocity through the
stalled tubes. Since the vapor was no longer in equilibrium with boiling EO it could
momentarily attain the 150o
C temperature of the reboiler steam supply.
3. The insides of the reboiler tubes had collected a thin film of EO polymer containing
percent-level amounts of catalytic iron oxides. This film had in numerous places
peeled away from the tube wall producing a catalytic surface of low heat capacity
and negligible effect on mass flow rate. EO vapor heating was aided by the absence of
liquid plus the small vapor velocity through the stalled tubes. These conditions led
to a rapid rate of film heating which encouraged a fast disproportionation reaction of
EO to predominate over slower polymerization reactions. The previously unknown
fast reaction between EO vapor and supported high surface area iron oxide led to a
hotspot and initiation of vapor decomposition. Once ignited the EO decomposition
flame spread rapidly through the column causing overpressurization.
9.6.2 Brittle Fracture of a Heat Exchanger
An olefin plant was being restarted after repair work had been completed. A leak devel-
oped on the inlet flange of one of the heat exchangers in the acetylene conversion preheat
system. To eliminate the leak, the control valve supplying feed to the conversion system
was shut off and the acetylene conversion preheat system was depressured. Despite the
fact that the feed control valve was given a signal to close, the valve allowed a small flow.
High liquid level in an upstream drum may have allowed liquid carryover which resulted
in extremely low temperature upon depressurization to atmospheric pressure.
The heat exchanger that developed the leak was equipped with bypass and block valves
to isolate the exchanger. After the leaking heat exchanger was bypassed, the acetylene
conversion system was repressured and placed back in service. Shortly thereafter, the first
exchanger in the feed stream to the acetylene converter system failed in a brittle manner,
releasing a large volume of flammable gas. The subsequent fire and explosion resulted in
two fatalities, seven serious burn cases, and major damage to the olefins unit.
The acetylene converter pre-heater failed as a result of inadequate lowtemperature resis-
tance during the low temperature excursion caused by depressuring the acetylene converter

system. The heat exchanger that failed was fabricated from ASTM A515 grade 70 car-
bon steel. After the accident, all process equipment in the plant which could potentially
operate at less than 200F was reviewed for suitable low-temperature toughness [116].
Ed. Note: It should have been recognized that upstream cryogenic conditions may have
a deleterious effect on downstream equipment during normal and abnormal operations.
9.6.3 Cold Box Explosion
Ethylene plants utilize a series of heat exchangers to transfer heat between a number of
low temperature plant streams and the plant refrigeration systems. This collection of
heat exchangers is known collectively as the ”cold box.” In one operating ethylene plant,
a heat exchanger in the cold box that handled a stream fed to the demethanizer column
required periodic heating and backflushing with methane to prevent excessive pressure
drop due to the accumulation of nitrogen-containing compounds.
During a plant upset which resulted in the shutdown of the plant refrigeration compressors,
the temperature of the cold box began to increase. During this temperature transient an
explosion occurred which destroyed the cold box and disabled the ethylene plant for about
5 months. An estimated 20 tons of hydrocarbon escaped. Fortunately, the hydrocarbon
did not ignite.
An investigation revealed that the explosion was caused by the accumulation and sub-
sequent violent decomposition of unstable organic compounds that formed at the low
temperatures inside the cold box. The unstable ”gums55 were found to contain nitro
and nitroso components on short hydrocarbon chains. The source of the nitrogen was
identified as nitrogen oxides (NOx) present in a feed stream from a catalytic cracking
unit. Operating upsets could have promoted unstable gums by permitting higher than
normal concentrations of 1, 3-butadiene and 1, 3-cyclopentadiene to enter the cold box.
To prevent NOx from entering the cold box, the feed stream from the catalytic cracking
unit was isolated from the ethylene plant [87].
9.7 Failure scenarios and design solutions
Table 9.1 presents information on equipment failure scenarios and associated design solu-
tions specific to heat transfer equipment.

9.7 Failure scenarios and design solutions 115
Figure 9.1. troubleshooting

9.8 Discussion
9.8.1 Use of Potential Design Solutions Table
To arrive at the optimal design solution for a given application, use Tables 9.1-9.4 in con-
junction with the design basis selection methodology presented earlier. Use of the design
solutions presented in the table should be combined with sound engineering judgment and
consideration of all relevant factors.

9.8 Discussion 117
9.8.2 Special Considerations
This section contains additional information on selected design solutions. The information
is organized and cross-referenced by the Operational Deviation Number in the table.
Leak/Rupture of the Heat Transfer Surface (1-3)
This common failure scenario may result from corrosion, thermal stresses, or mechanical
stresses of heat exchanger internals. The leak/rupture of tubes leads to contamination or
overpressure of the low-pressure side. Failure to maintain separation between heat transfer
and process fluids may lead to violent reaction in the heat transfer equipment or in the
downstream processing equipment. To make the heat transfer process inherently safer,
designers must look at possible interactions between heating/cooling fluids and process
fluids.
For relatively low-pressure equipment (<1000 psig), a complete failure of tubes may not
be a credible overpressure scenario if the design pressure of the low-pressure side and
associated equipment is greater than two-thirds of the design pressure of the high- pressure
side [2], or if the geometry of the tube layout is such that a complete break is not physically
possible. For high-pressure equipment (> 1000 psig), however, a complete failure should
be considered credible, regardless of pressure differential.
Double tube sheets or seal welding may be used for heat exchangers handling toxic chem-
icals. For heat transfer problems involving highly reactive/ hazardous materials, a triple-
wall heat exchanger may be used. This type of heat exchanger consists of three chambers
and uses a neutral material to transfer heat between two highly reactive fluids. Alter-
natively two heat exchangers can be used with circulation of the neutral fluid between
them.
There are known cases of cooling tower fires that have resulted from contamination of
cooling water with hydrocarbons attributable to tube leakage. Gas detectors and separa-
tors may be installed on the cooling water return lines, or in the cooling tower exhaust
(air) stream.
Thermal stresses can be reduced by limiting the temperature differences be-
tween the inlet and outlet streams. In addition, alternate flow arrangements may be
used to avoid high thermal stresses. Thermal cycling of heat transfer equipment should
be kept to a minimum to reduce the likelihood of leaks and ruptures.
Fouling, or Accumulation of Noncondensable Gases (5)
It is desirable to design heat exchangers to resist fouling. Sufficient tube side velocity may
reduce fouling. However, higher tube side velocities may also lead to erosion problems.
In some cases fouling will cause higher tube wall temperatures, leading to overheating of
reactive materials, loss of tube strength, or excessive differential thermal expansion.
Accumulation of noncondensable gases can result in loss of heat transfer capability. Heat
exchangers in condensing service may need a vent nozzle, or other means of removing
noncondensable gases from the system.
External Fire (9)
Emergency relief devices are often sized for external fire. Heat transfer equipment, such
as air coolers, present a unique challenge when it comes to sizing relief devices. These
exchangers are designed with large heat transfer areas. This large surface area may result
in very large heat input in case of external fire. Indeed, it may not be practical to install
a relief device sized for external fire case due to large relief area requirements. Other

mitigation measures, such as siting outside the potential fire zone or diking with sloped
drainage, may be used to reduce the likelihood and magnitude of external fire impinging
on the heat exchanger. Alternative heat exchanger designs may also be used to reduce
the surface area presented to an external fire.
9.9 Troubleshooting Examples
9.9.1 Shell side temperature uncontrolled
55 Co
Control
vlave
30 Co
Organic
55-62 Co
uncontrolled
125 Co
Water
67 Co
Control
vlave
30 Co
Organic
55-62 Co
controlled
70 Co
Water
Bypass
Symptom: Shellside outlet
temperaturee cannot be
controlled within desired
range (55-62 C) byo
controlling flow of 125 Co
water to tubes. The heat
exchanger is 4 tube pass.
Diagnosis: Heat exchanger is
considerablyo versized for the
duty (because of an alternative
service). Temperature correction
factors F for LMTD fluctuate
widely with small changes in
tube side flow
Cure: Tube side water
temperature reduced to 70oC
and control valve removed.
Control valve is installed
in new shellside bypass
line
Figure 9.5. Shell side temperature uncontrolled
9.9.2 Shell assumed banana-shape
Symptom: Shell assumed
banana shape and piping
connections leaked. leakage
between tube and shell side
Diagnosis: vertically cut baffle
and inlets and outlets of top shell
side, caused stratification of
gases at top of shell. Poor
distribution of hot gases lead
to unequal expansionof tubes
Cure: increase the number of baffles
from two to three; weld baffles in the
shell; install sealing strips at edges of
bundle; installed three concentric cones
in tube side inlet; install vapor belt - for
shellside inlet nozzle; change baffles
from vertical to horizontal cut.
belows joint
487 Co
200 Co
560 Co
600 Co
Figure 9.6. Shell assumed banana-shape

9.9 Troubleshooting Examples 119
9.9.3 Steam condenser performing below design capacity
Symptom: Air cooled steam
condensor performing below
design capacity.
Diagnosis: Careful measurement
tube levels discloded that tubes
sloped 1/4 inch in wrong direction
(rising toward condensate end)
Cure: Raise inlet end to obtain 2 inch slope
toward condensate outlet
Steam
Vent
ondensate
Figure 9.7. Steam condenser performing below design capacity
9.9.4 Steam heat exchanger flooded
When a heat exchanger ”stalls,” condensate floods the steam space and causes a variety
of problems within the exchanger:
Figure 9.8. Conventional motor driven condensate pump system
• Control hunting: As condensate backs up in the exchanger, the heat transfer rate to
the process is greatly reduced. The control valve opens wide enough to allow flow
into the exchanger. As condensate drains out, the steam space is now greater and
the steam pressure increases. The process overheats, the control valve closes down,
and the cycle repeats.
• Temperature shock: Condensate backed up inside the steam space cools the tubes
that carry the process fluid. When this sub-cooled condensate is suddenly replaced
by hot steam due to poor steam trap operations, the expansion and contraction of
the tubes stress the tube joints. Constantly repeating this cycle causes premature
failure.
• Corrosion from:
1. Flooding - A flooded heat exchanger will permit the oxygen to dissolve, as well
as carbon dioxide and other gases found in the steam. Because the condensate

120 10 Unresolved problems in the heat exchangers design
is often sub-cooled due to the time it is in the exchanger, these gases are more
readily dissolved. Together the cool condensate and dissolved gases are ex-
tremely corrosive and will tend to decrease the efficiency of the heat exchanger
and reduce the heat transfer through the tubes.
2. Steam collapse - Under very low loads with the steam valve closed, the steam
volume collapses to smaller volume condensate, inducing a vacuum. When
the vacuum breaker opens, atmospheric air and condensate mix inside the
exchanger, increasing the possibility of corrosion of the tubes, shells, tube
sheet and tube supports.
3. Freezing - Steam/air coils cannot afford poor condensate drainage, especially
if the coil experiences air below freezing temperature. Condensate backed up
inside the coil will freeze, often within seconds, depending on the air temper-
ature. A low temperature detection thermostat is recommended on the coil
leaving side to sense freezing conditions. As we previously explained, the only
way to avoid ”stall” is to eliminate back pressure on the steam trap. There are
a number of options available for designing a system that greatly reduces the
risk of ”stall.” The following are two such options:
• Install the heat exchanger in a position so that the condensate freely drains by
gravity to the condensate return line. In many cases this is not possible because
of existing piping around the area in which the heat exchanger is needed (e.g., the
heat exchanger is installed at a level lower than the condensate return tank).
• Use an electric or pressure driven condensate pump package installed below the
steam trap to pump condensate back to the boiler.
In actual practice, the first option may not be possible, and so the use of electric or
pressure driven pumps to return condensate to the boiler room should be considered.
10 Unresolved problems in the heat exchangers de-
sign
1. Accurate data on the thermodynamic properties: These are needed for both pure
fluid and mixtures in single phase and two phase system under extremes conditions.
It would be best if more predictable methods could be obtained
2. fouling (predictive method not available)
3. flow induced vibration (prediction)
4. two phase flow (flow regime)
5. boiling of mixture (heat transfer coefficient)
6. turbulence (better understanding)
10.1 Future trend
1. Stepwise calculation of overall heat transfer coefficient instead of assumption
2. Thermodynamic properties from built-in subroutines
3. workshops fabrication drawings.
4. better transportation facilities for the shell of heat exchanger.
5. computer design code

Bibliography 121
Bibliography
[1] Adunka, A.: Meßunsicherheiten ”Theorie und Praxis” 2.Auflage, VULKAN-Verlag,
Essen, Germany, 2000
[2] API RP 521. Guide for Pressure Relieving and Depressuring Systems. Washington
D.C.: American Petroleum Institute. 1993
[3] Baehr, H.D.; Stephan, K.: Waerme-und Stoffübertragung. (2ed), Springer, Berlin,
1996
[4] Baker, O.: Design of pipe lines for simultaneous flow of oil and gas. Oil and Gas J.,
July 26. Bao, Z.Y., 1954
[5] Bandel, J.: Druckverlust und Wärmeübergang bei der Verdampfung siedender
Kältemilttel im durchströmten waagerechten Rohr. Dissertation, Universität Karl-
sruhe, Germany, 1973
[6] Baumann, P.: Zur Thermohydraulik von Gas/Dampf-flüßigkeitsgemischen in hori-
zontalen Rohren. Dissertation, Universitat Karlsruhe, Germany, 1993
[7] Beattice, D.R.H.; Lawther, K.R.: An examination of wall temperature drop phe-
nomenon during approach to flow boiling crisis. Proc. 8th Int. Heat Transfer Con-
ference, San Francisxo, USA, 1986
[8] Bennet, D.L.; Chen, J.C.: Forced convective boiling in vertical tubes for saturated
pure components and binary mixtures. AICHE Journal, Vol. 26, 1980, 454-461
[9] Bertoletti, S.; Lombari, C.; Silvestri, M.: Heat transfer to steam-water mixtures,
C.I.S.E., Report R-78, 1964
[10] Boissieux, X.; Heikel, M.R.; Johns, R.A.: Two-phase heat transfer coefficients of
three HFC refrigerants inside a horizontal smooth tube, part I: Evaporation. Int. J.
of Refrig., Vol. 23, 2000, 269-283
[11] Bonilla, C.F.; Perry, C.W.: Heat transimission to boiling mixtures. Am. Inst. Chem.
Eng. J., Vol. 37, 1941, 685-705
[12] Bonn, W.: Wärmeübergang und Druckverlust bei der Verdampfung von Stickstoff
und Argon in durchströmten horizontalen Rohr sowie Betrachtungen über die tan-
gentialle Wärmeleitung und die maximal möglische Flüssigkeitsüberhitzung. Disser-
tation, Universität Karlsruhe (TH), Germany, 1980
[13] Bonn, W.; Iwicki, J.; Krebs, R.; Steiner, D.; Schluender, E. U.: Über die Auswirkung
der Ungleichverteilung des Wärmeübergangs am Rohrumfang bei der Verdampfung
im durchströmten waagerechten Rohr. Wärme-und Stoffübertragung, 1980, 265-274
[14] Branan, C. R.: Process Engineer’s Pocket Handbook, Vol. 1, Gulf Publishing Com-
pany, 1976.
[15] Butterworth, D.: “Condensors: basic heat transfer and fluids flow”, in Kakac, S.;
Bergles, A.E.; Mayinger, F. (eds): Heat exchangers. Hemisphere publishing Corp.,
New York, USA, 1981, 289-314
[16] Calus, W.F.; di Montegnacco, A.; Denning, R.K.: Heat transfer in a natural circu-
lation single tube reboiler, part II: Binary liquid mixtures. Chem. Eng. J., Vol. 6,
1973, 251-264
[17] Calus, W.F.; Leonidopoulos, D.J.:Pool boiling-binary liquid mixtures. Int. J. of Heat
and Mass Transfer, Vol. 17, 1974, 249-256

122 Bibliography
[18] Carey, V.P.: Liquid-vapor-phase-change phenomina: An introduction to the ther-
mophysics of vaporization and condensation processes in heat transfer equipment.
Tayler and Francis, Bristol, UK, 1992
[19] Chen, J.C.: Correlation for boiling heat transfer to saturated fluids in convective
flow. I and EC Process Design and Development, Vol. 5, 1966, 322-29
[20] Chen, M.M.: An analytical study of laminar film condensation : Part 1-Flat plates
and part 2-Multiple horizontal tubes. J. of Heat Transfer, Vol. 83, 1961, 48-54, 55-60
[21] Chen, L.-T.; Huang, F.; Rong-Fung, M.: Boiling heat transfer to R22/DMF mix-
tures. Int. Comm. Heat and Mass Transfer, Vol. 12, 1985, 541-549
[22] Chisholm, D.: Pressure gradients due to friction during the flow of evaporating two
phase mixtures in smooth tubes and channels. Int. J. of Heat and Mass Transfer,
Vol. 16, 1973, 374-358
[23] Churchil, S.W.; Chu, H.H.S.: Correlation equations for laminar and turbulent free
convection from a vertical plate. Int. J. of Heat and Mass Transfer, Vol. 18, 1975,
1323-1329
[24] Chwala, J.M.: Wärmeübergang und Druckabfall in waagerechten Rohren bei der
Strömung von verdampfenden Kältemitteln. VDI-Forschungsheft Nr. 523, VDI-
Verlag GmbH, Düsseldorf, Germany, 1967
[25] Collier, J.G.; Thome, J.R. (ed): Convective boiling and condensation. 3rd ed.,
Oxford University Press, Oxford, UK, 1994
[26] Collier, J.G.: Gas-liquid flow, in: heat exchanger design handbook. Vol. 2. Hemi-
sphere, Washington, DC, USA, 1983
[27] Cooper, M.G.: Saturation nucleate pool boiling, A simple correlation. 1st UK Na-
tional Conference on Heat Transfer, Vol. 2, 1984, 785-793
[28] Costigan, G.; Frankum, D.P.; Wadekar, V.V.: Flow boiling measurements on pen-
tane, iso-octane and pentane/iso-octane mixtures. Proc. 10th Int. Heat Tranafer
Coeference,Brighton, UK, Vol. 7, 1994,431-436
[29] Coulson and Richardson’s chemical engineering volume 6 (3rd ed). Butterworth and
heinemann, Oxford, 1999
[30] Dembi, et al.: in Aumann, L.: Verfahren zur Berechnung lokaler
Wärmeübergangkoeffizienten von Zweiphasenströmungen in Rohren. Studienarbeit,
Institut für Thermodynamik, Universität Hannover, Germany, 1998.
[31] Der Deutsche Kalibrierdienst (DKD): Ermittlung von Meßunsicherheiten (DKD-3).
Physikalisch-Technische-Bundesanstlt, Braunschweig, 1991
[32] Didion, D.A.; Bivens, D.B.: Role of refrigerant mixtures as alternative to CFCs.
Int. J. of Refrig., Vol. 13, 1990, 163-175
[33] Dittus, F.W.; Boelter, L.M.K.: University of California (Berkeley) Publications on
Engineering, Vol. 2, University of California, Berkeley, CA, 1930, 443
[34] Domanski, P.A.; Didion, D. A.: Thermodynamic evaluation of R-22 alternative
refrigerants and refrigerant mixtures. ASHRAE Trans., Vol. 99, Part 2, 1993, 636-
648

Bibliography 123
[35] Eckels, S.J.; Doerr, T.M.; Pate, M.B.: In-tube heat transfer and pressure drop of
R-134a and Ester lubricant mixtures in a smooth tube and a micro-fin tube: Part
I-Evaporation. ASHRAE Trans., Vol. 100, 1994, 265-282
[36] Eckels, S.J.; Doerr, T.M.; Pate, M.B.: Heat transfer coefficient and pressure drop
for R-134a and an Ester lubricant mixture in a smooth tube and micro fin tube.
ASHRAE Trans., Vol. 104, 1998, 366-375
[37] Eckels, S.J.; Doerr, T.M.; Pate, M.B.: A comparison of the heat transfer and
pressure drop for R-134a and lubricant mixture in different diameter smooth tubes
and nicro fin tubes. ASHRAE Trans., Vol. 104, Part A, 1998, 376-386
[38] Eckels, S.J.; Pate, M.B.: An experimental comparison of evaporation and conden-
sation heat transfer coefficients for HFC-134a and CFC-12. Int. J. Refrig., Vol. 14,
March, 1991, 70-78
[39] Emhart (1969), Heat Exchanger Tube-to-Tubesheet Joints, ASME Pap. 69-WA/HT-
47.
[40] Fenghour et al.: Transport properties of ammonia. Int. J. Chem. Phys. Ref. Data,
Vol. 24, No. 5, 1995, 1649-67
[41] Fletcher, D.F.; Haynes, B.S.: An experimental study of gas-liquid flow in a narrow
conduit. Int. J. of Heat and Hass Transfer, Vol. 43, 2000, 2313-24
[42] Friedel, L.: Momentum exchange and pressure drop in two-phase flow. Proceeding
NATO Advanced Study Institute, Istanbul, Turkey, 16-17 August, Vol. I, 1976,
239-312
[43] Fuchs, P.H.: Influence of the tube material and external heat load on heat transfer
coefficient with separated flow in evaporator. Proc. of the 16th Int. Congr. of Refrig.,
Venedig, Vol. 2, 1987, 503-506
[44] Fujita, Y.; Bai, Q.: Bubble dynamics and heat transfer in mixture boiling. Pro-
ceeding of the 12th Int. Heat Transfer Conference, Gernoble, France, Vol. 1, 2002,
93-104
[45] Fujii, T.; Uehara, H.; Kurata, C.: Laminar filmwise condensation of flowing vapor
on a horizontal cylinder. Int. J. of Heat and Mass Transfer, Vol. 15, 1972, 235-246
[46] Gnielinski, V.: New equation for heat and mass transfer in turbulent pipe and
channel flow. Int. J. Chemical Engineering, Vol. 16, 1976, 359-368
[47] Gorenflo, D.; Bieling, V.: Heat transfer at pool boiling of mixtures with R22 and
R115. XVII Int. Symp. on Heat and Mass Transfer in Cryoeng and Refrig., 1986,
243-257
[48] Goebel, O.: Modeling of two phase stratified and annular flow in heated horitontal
tubes. In: Mayinger, F.; Lehner, M.: Convective flow and pool boiling. Taylor and
Francis, Philadelphia, USA, 1999
[49] Granryd, E.: Hydrocarbons as refrigerants- an overview. Internationl J. of Refrig.,
Vol. 24, 2001, 15-24
[50] Gropp, U.:Wärme- und Stoffübertragung bei Oberflächen Verdampfung und beim
Blasensieden eines Binäres Gemisches am Riesenfilm. Dissertation, Universität Karl-
sruhe (TU), Germany, 1989

124 Bibliography
[51] Gross, U.; Song, Y.W.; Hahne, E.: Thermal conductivity of the new refrigerants
R134a, R152a and R123 measured by transient hot-wire method. Int. J. of Ther-
modynamics, Vol. 13, No. 6, 1992, 957-983
[52] Gungor, K.E.; Winterton, R.H.S.: General correlation for flow boiling in tubes and
annuli. Int. J. of Heat and Mass Transfer, Vol. 26, 1986, 351-358
[53] Hambraeus, K.: Heat transfer coefficient during two-phase flow boiling of HFC-134a.
Int. J. Refrig, Vol. 14, No. Nov.,1991, 357-362
[54] Hashizume, K.: Flow pattern and void fraction of refrigerant two-phase flow in a
horizontal pipe. Bulletin of the JSME, Vol. 26, No. 219, 1983, 1597-1602
[55] Herbst, K.E.; Fechner, A.; Köhler, W.; Goebel, O.; Oberle, B.: Heat transfer behav-
ior of an absorber tube with direct steam generation by water injection. In: Chen,
J.C. (ed). Convective flow boiling. Taylor & Francis, 1996
[56] Hewitt, G.F.; Kearcy, H.A.; Lacey, P.M.C.: Burn-out and nucleation in climbing
film flow. Int. J. of Heat and Mass Transfer, Vol. 8, 1965, 793-814
[57] Hihara, E.; Tanida, K.; Saito, T.: Forced convection boiling experiments of binary
mixtures. Proc. of the ASME-JSME Thermal Engineering joint Conf., Honolulu,
Publ. by ASME, Vol 5, 1989, 119-126
[58] Holcomb, C.D.; Magee, J,W.; Scott, J.L.; Outcalt, S.L.; Haynes, W.M.: Selected
thermodynamic properties for mixtures of R-32 (Difluoromethane), R125 (pentaflu-
oroethane), R134a (1,1,1,2-tetrafluoroethane), R143a (1,1,1-trifluoroethane),
R41(fluoroethane), propane (R290), and R744 (carbon dioxide). NIST Technical
Note No. 1397, U.S. Dept. Commerce, Washington, 1997
[59] ICI: Physikalische Eigenschaften von Klea 134a SI-Einheiten. Mitteilung der Fa.
Deutsche ICI GmbH, Frankfurt, Germany, 1993
[60] Jain, V.K.; Dhar, P.L.: Studies on flow boiling of mixture of refrigerants R12 and
R13 inside a horizontal tube. Proc. XVI, Int. Congr. Refrig., Paris, 1983
[61] Jallouk, P.A.: Two phase flow pressure drop and heat transfer characteristics of
refrigerants in vertical tube. PhD-thesis, University of Tennessee, 1974
[62] Jensen, M. K.; Bensler, H.P.: Saturated forced convection boiling heat transfer with
twisted tape inserts. ASME J. Heat Transfer, Vol. 108, 1986, 93-99
[63] Jung, D.S.: Horizontal flow boiling heat transfer using refrigerant mixture. PhD-
thesis, University of Maryland, 1988
[64] Jung, D.S.; McLinden, M.; Radermacher, R.; Didion, D.: Horizontal flow boiling
heat transfer experiments with a mixture of R22/R114. Int. J. of Heat and Mass
Transfer, Vol. 32, 1989, 131-145
[65] Jung, D.S.; Radermacher, R.: Prediction of evaporation heat transfer coefficient and
pressure drop of refrigerant mixtures. Int. J. Refrig., Vol. 16, No. 5, 1993, 330-338
[66] Jung, D.; Song, Y.; Park, B.: Performance des melanges de frigorigenes utilises pour
remplacer le HCFC22. Int. J. of Refrig., Vol. 23, 2000, 466-474
[67] Kabelac, S.; de Buhr, H-J.: Flow boiling of ammonia in a plain and low finned
horizontal tube. Int. J. of Refrig., Vol. 24, 2001, 41-50

Bibliography 125
[68] Kabelac, S.; de Buhr, H.-J.: Flow boiling of ammonia and ammonia -oil mixtures
in a plain and a low finned horizontal tube. Proc. 34th National Heat Transfer
Conference, 2000,1-8
[69] Kabelac, S.; Rabah, R.: Flow boiling of R134a-propane mixtures: -experiments and
CML modeling-. Proc. of the 12th. Int. Heat Tranasfer Conference, Grenoble, Vol.
4, 2002, 525-530
[70] Kandlikar S.G.: A general correlation for two-phase flow boiling heat transfer inside
horizontal and vertical tubes. Int. J. Heat and Mass Transfer, Vol. 112, 1990, 219-228
[71] Kandlikar, S.G.: Boiling heat transfer with binary mixtures: Part II- Flow boiling
in plain tubes. Transaction of ASME, Vol. 120, 1998, 388-394
[72] Kandlikar, S.G.: Boiling Heat transfer with binary mixtures: Part I- A theoretical
model for pool boiling. Transaction of ASME, vol. 120, 1998, 380-387
[73] Kandlikar, S.G.; Shoji, M. ; Dhir, V. (ed): Hand Book of Phase Change: Boiling
and Condensation. Taylor & Francis, Philadelphia, 1999
[74] Kapitza, P.L.: Wave flow of thin viscous fluids. Zh. Eksp. Teoret. Fitz., Vol. 18,
1948, 1-1
[75] Kattan, N.; Thome, J.R.; Favrat, D.: Flow boiling in horizontal tubes. Part 1-
Developmment of adiabatic two phase flow pattern map. Trans. of ASME, Vol. 120,
1998, 140-147
[76] Kattan, N.; Thome, J.R.; Favrat D.: Flow boiling in horizontal tubes. Part 2-New
heat transfer data for five refrigerants. ASHRAE Trans., Vol. 120, 1998, 148-155.
[77] Kattan, N.; Thome, J.R.; Favrat D.: Flow boiling in horizontal tubes. Part 3-
Development of a new heat transfer model based on flow pattern map. ASHRAE
Trans., Vol. 120, 1998, 156-165
[78] Kenning, D. B.R.; Cooper, M. G.: Saturated flow boiling of water in vertical tubes.
Int. J. of Heat and Mass Transfer, Vol. 31, 1988, 455-458.
[79] Kern, D. Q.: Process Heat Transfer, McGraw-Hill, 1950.
[80] Khanpara, J.C.: Augmentation of in tube evaporation and condensation with micro-
fin tubes. PhD-thesis, Iowa State University, Ames, IA, 1986
[81] Kim, M.S.; Mulroy, W.J; Didon, D.A.: Performance evaluation of two azeotropic
refrigerant mixtures of HFC-134a with R-290 (propane) and R-600 (isobutane).
Trans. of the ASME, Vol. 116, June, 1994, 148-154
[82] Kleemiß, M.: Thermodynamische Eigenschaften zweir ternärer Kältemittelgemi-
sche. Dissertation, Universität Hannover, Germany,1997
[83] Kleiber, M.: Vapor-liquid equilibria of binary refrigerant mixtures containing propy-
lene or R134a. Fluid Phase Equil. Vol. 92, 1994, 149-194
[84] Klimenko, in: Aumann, L.: Verfahren zur Berechnung lokaler Wärmeübergang-
koeffizient von Zweiphasenströmungen in Rohren. Studienarbeit, Institut für Ther-
modynamik, Universität Hannover, Germany, 1998.
[85] Koh, C.Y.; Sparrow, E.M.; Harnett, J.P.: The two phase boundary layer in laminar
film condensation. Int. J. Heat Mass Transfer, Vol. 2, 1961, 69-82.

126 Bibliography
[86] Köhler, W.; Kefer, V.; Kastner, W.: Bestimmung des Wärmeübergangs im un-
benetzen Teil von Verdampferrohren unter Berücksichtigung zweidimensionaler
Wärmeleitung in Wandmaterial. Wärme-Stoffübertragung, Vol. 21, 1987, 133-138
[87] Kohler, J.1991. Cold Box Explosion at Shell Steam Cracker in Berre, France. Paper
presented at AIChE Spring National Meeting, Houston, Texas.
[88] Kon’kov, A.S.: Experimental study of the conditions under which heat exchanger
deteriorates when a steam-water mixture flows in heated tubes. Teploenergetika,
Vol. 13, No. 12, 1965, 77
[89] Lee, W.C.; Rabbar, S.; Rose, J.W.: Film condensation of refrigerant R113 and
ethanediol on a horizontal tube- effect of vapour velocity. ASME J. Heat Transfer,
Vol. 106, 1984, 524-530
[90] Liu, Z.; Winterton, R.H.S,: A general correlation for saturated and subcooled flow
boiling in tubes and annuli based on a nucleate pool boiling equation. Int. J. of Heat
and Mass Transfer, Vol. 34, No. 8, 1990, 2759-2766
[91] Lockhart, R.W.; Martinelli, R.C.: Proposed correaltion for data for two-phase, two-
component flow in pipes. Chem. Eng. Prog. Vol. 45, 1949, 39-48
[92] Lucas, K.; Luckas, M.: Berechnungsmethoden für Stoffeigenschaften. VDI-
Wärmeatlas. 8. Auflage, Springer, Berlin, 1997
[93] Ludwig, E. E.: Applied Process Design for Chemical and Petrochemical Plants, Vol.
3, 3rd edition, Gulf Publishing Company, 1998.
[94] Mathur, G.D.: Heat transfer coefficient for propane (R-290), isobutane (R-600a),
and 50/50 mixture of propane and isobutane. ASHRAE Transaction, Vol. 104, No.
2, 1998, 1159-1172
[95] McAdams, W.H.; Woods, W. K.; Heroman, L.C.: Vaporization inside horizontal
tube-II, Benzene oil mixtures. Trans. ASME, Vol. 64, 1972, 193-193
[96] McLinden, M.O.: Measurement and formulation of the thermodynamic properties
of refrigerants R134a, R123. ASHRAE trans., Vol. 95, No. 2, 1989, 79-103
[97] Melin, P.; Vamling, L.: Flow boiling heat transfer and pressure drop for HFC-134a
in a horizontal tube. Proc. of meetings of commissions B1, B2, E1, E2, Padoua,
1994, 575-582
[98] Memory, S.B.; Rose, J.W.: Free convection laminar film condensation on a horizon-
tal tube with variable wall temperature. Int. J. of Heat and Mass Transfer. Vol. 34,
No. 11, 1991, 2775-2778.
[99] Moldover, M.R.; Trusler, J.P.M.; Edwards, T.J.; Mehl, J.B.; Davis, R.S.: Mea-
surement of the universal gas consatnat R using a spherical acoustic resonator. J.
Research NBS, Vol. 93, No. 2, 1988, 85
[100] Morrison, G; McLinden, M.O.: Azeotropy in refrigerant mixture. Rev. Int. Froid.,
Vol. 16, No. 2, 1993, 129-137
[101] Müller-Steinhagen, H.: Wärmeübergang und Fouling beim Strömungssieden von
Argon und Stickstoff im horizontalen Rohr. Fortschritt-Berichte VDI- Reihe 6 Nr.
143, VDI-Verlag GmbH, Düsseldorf, Germany, 1991
[102] Müller-Steinhagen, H.; Schlünder, E.U.: Über den Einfluß des Wärmeleitvermögens
der Rohrwand auf den umfangsmittelen Wärmeübergangs-koeffizienten beim Sieden
in horizontalen Verdampferrohren. Chem. Eng. Process, Vol. 18, 1984, 303-316

Bibliography 127
[103] Murata, K.; Hashizume, K.: Forced convection boiling of non-azeotropic refrigerant
mixtures inside tubes. ASME J. of Heat Transfer, Vol. 115, 1993, 680-689
[104] Murata, K.; Hashizume, K.: An experimental invistigation on forced convection
boiling of nonazeotropic refrigerant mixtures. Heat transfer Jpn. Res., Vol. 19, No.
2, 1990, 95-109
[105] Niederkrüger, M.: Strömungssieden von reinen Stoffen und binären zeotropen
Gemischen im waagerechten Rohr bei mittleren und hohen Drücken. Fortschritt-
Berichte VDI- Reihe 3, Nr. 245, VDI-Verlag GmbH, Düsseldorf, Germany, 1991
[106] Niederkrüger, M.; Steiner, D.; Schluender, E.-U.: Horizotal flow boiling experiments
of saturated pure components and mixtures of R846-R12 at high pressure. Int. J.
Refrig., Vol. 15, No. 1, 1992, 48-58
[107] Niederkrüger, M.; Steiner, D.: Flow boiling heat transfer to saturated pure compo-
nents and non-azeotropic mixtures in a horizontal tube. Chemical Engineering and
Processing, Vol. 33, 1994, 261-274
[108] NIST: REFRPOP. National institute for science and technology. Washington, DC,
1995
[109] Nusselt, W.: Die Oberflächenkondensation des Wasserdampfes. VDI-Z. 60, 1916,
541-546, 569-575
[110] Okubo, T.; Hasuo, T.; Nagashima, H.A.: Measurement of viscosity of HFC 134a in
the temperature range 213-423 K and at pressure up to 30 MPa. Int. J. of Thermo-
dynamics, Vol. 13, No. 6, 1992, 931-942
[111] Palen, J.W.: Falling film evaporation of wide-boiling-range mixtures inside a vertical
tube. PhD-thesis, Lehigh Univ., 1988
[112] Perry, R.H.; Green, D.W.: Perry’s chemical engineers’ hand book. 16th (ed), Mc-
Graw Hill, New York, 1984
[113] Perry, R. H., Green, D. W. (Eds.): Perry’s Chemical Engineers’ Handbook, 7th
edition, McGraw-Hill, 1997 , Section 11.
[114] Petukhov, B.S.; Popov, V.N.: Theoretical calculation of the heat exchanger and fric-
tional resistance in turbulent flow in tubes of an incompressible fluid with variable
physical properties. teplofiz. Vysok. temperatur (High temperature heat physics)
Vol. 1, No. 1, 1963
[115] Price, J. H. 1989. Personal communication to T.W. Carmody, Director CCPS.
[116] Price, J. H. 1989. Personal communication to T.W. Carmody, Director CCPS.
[117] Rauhani, S.Z.: Subcooled void fraction. AB Atomenergie (Sweden) Report AE-
RTV 841, 1969
[118] Reid, R.C.; Prausnitz, J. M.; Poling, B. E.: The properties of gases and liquids.
McGraw-Hill, New York, 1987
[119] Reynolds, Tube Welding for Conventional and Nuclear Power Plant Heat Exchang-
ers, ASME Pap. 69-WA/HT-24, 1969
[120] Rohlin, P.: Heat transfer coefficient of zeotropic mixtures and their pure compo-
nents in horizontal flow boiling -an experimental study. Proceedings of 1997 ASME
International Mechanical Engineering Congress and Exposition, Dallas, USA, 1997

128 Bibliography
[121] Rohsenow, W.M.; Hartnett, J.P.; Cho, Y.I.: Hand Book of Heat Transfer. McGraw-
Hill, New York, 1998
[122] Rose, J.W.: Effect of pressure gradient in forced convection film condensation on a
horizontal tube. Int. J. of Heat and Mass Transfer, Vol. 27, No. 1, 1984, 39-47
[123] Rose, J.W.: Condensation heat transfer fundamental. Trans. IchemE, Vol.76, part
A, 1998, 143-152
[124] Rose, J.W.: Fundamental of condensation heat transfer: Laminar film condensation.
JSME Int. J. of Heat and Mass Transfer, Series II, Vol. 31, No. 3, 1988, 357-375
[125] Ross, H.D.; Radermacher, R.; Di Marzo, M.; Didion, D.: Horizontal flow boiling of
pure and mixed refrigerants. Int. J. Heat Mass Transfer, Vol. 30, 1987, 979-992
[126] Schlünder, E.U.: Über den Wärmeübergang bei der Blasenverdampfung von Gemis-
chen. vt Verfahrenstechnik, Vol. 16, No. 9, 1982, 692-698
[127] Schluender, E. U. (Ed.): VDI Heat Atlas, Woodhead Publishing, Limited, 1993,
Chapter Cc.
[128] Schmidt, H.: Beitrag zum Verständnis des Wärmeübergangs im horizontalen
Verdampferrohr. Fortschritt-Berichte VDI- Reihe 19, No. 6, VDI-Verlag GmbH,
Düsseldorf, Germany, 1986
[129] Schrock, V.E.; Grossman, L. M.: Forced convection in tubes. Nuclear Science and
Engineering, Vol. 12, 1962, 474-481
[130] Shah, M.M.: A new correlation for heat transfer during boiling flow through pipes.
ASHREE Trans., Vol. 82, 1976, 66-86
[131] Shah, M.M.: Prediction of heat transfer during boiling of cryogenic fluids flowing
in tubes. Cryogenics, Vol. 5, 1984, 231-236.
[132] Shah, M.M.: Chart correlation for saturated boiling heat transfer: Equation and
further studies. ASHREE Trans. Vol. 88, No. 1, 1982, 80-89
[133] Shao, D.W.; Granryd, E.: Experimental and theoretical study on flow condensation
with non-azeotropic refrigerant mixtures of R32/R134a. Int. J. Refrig. Vol. 21, No.
3, 1998, 230-246
[134] Shekriladze, I.G.: Gomelauri, V.I.: Theoretical study of laminar film condensation
of flowing vapor. Int. J. of Heat and Mass Transfer. Vol.9, 1966, 581-1
[135] Shin, J.Y.; Kim, M.S.; Ro, S.T.: Experimental study on forced convective boiling
heat transfer of pure and refrigerant mixtures in a horizontal tube. Int. J. of Refrig.,
Vol. 20, No. 4, 1997, 267-275
[136] Shoji, M.: Boiling simulator- a simple theoretical model of boiling-. Proc. of the 3rd
Int. Conf. on Multiphase Flow, Lyon, France, June 1998, 8-12
[137] Singal, L.C.; Sharma, C.P.; Varma, H.K.: Heat transfer correlation for the forced
convection boiling of R12/R13 mixture. Int. J. Refrig., Vol. 7, No. 5, 1984, 278-284
[138] Sparrow, E.M.; Gregg, J.L.: A boundray-layer treatment of laminar film condensa-
tion. J. Heat Transfer, No. 81C, 1959, 13-18,
[139] Spindler, K.: Flow boiling. Proc. Int. Heat Transfer Conference, Brighton, UK, Vol.
1, 1994, 349-369

Bibliography 129
[140] Steiner, D.: Strömungssieden Gesättiger Flüssigkeiten. in: VDI, VDI-GVC: VDI-
Wärmeatlas. 8. Aufl., Springer-Verlag, Berlin, 1997, Abschnitt, Dbb
[141] Stephan, K.: Two-phase heat exchange for new refrigerants and their mixtures. Int.
J. Refrig., Vol. 18, No. 3, 1995, 198-209
[142] Stephan, K.; Korner, M.: Berechnung des Warmeübergangs verdampfender binärer
Flüßigkeitsgemische. Chem. Ing. Tech., Vol. 2, 1969, 161-169
[143] Stephan, K.; Preusser, P.: Heat transfer and critical heat flux in pool boiling of
binary and ternary mixtures. German Chem. engng, Vol. 2, 1979, 161-169
[144] Storek, H.; Brauer, H.: Reibungsdruckverlust der adiabaten Gas/Flüßigkeits-
stömung in horizontalen und vertikalen Rohren. VDI-Forschungsheft, Nr. 599, VDI-
Verlag GmbH, Düsseldorf, Germany, 1980
[145] Taitel, Y.; Dukler, A.E.: A model for predicting flow regime transitions in horizontal
and near horizontal Gas-Liquid flow. AICHE J., Vol. 22, No. 2, 1985, 43-55
[146] In: Heat exchanger design handbook, New York: Hemisphere, Volo. 1, section 1.5
[147] Tubular Exchanger Manufacturers Association, ”Standards of the Tubular Ex-
changer Manufacturers Association,” 7th ed., TEMA, New York (1988).
[148] Thome, J. R.: Prediction of binary mixture boiling heat transfer coefficient using
only phase equilibrium data. Int. J. of Heat and Mass Transfer, Vol. 26, 1983, 965-
974
[149] Thome, J. R.: Enhanced boiling of mixture. Chemical Engineering Service, Vol. 42,
No. 8, 1986, 1909-1917
[150] Tillner-Roth, R.: Die thermodynamischen Eigenschaften von R152a, R134a und
ihren Gemischen-Messungen und Fundamental Gleichungen. Dissertation, Univer-
sität Hannover; Germany, 1993
[151] Tillner-Roth, R.: Fundamental equations of state. Shaker-Verlag, Aachen, Germany,
1998
[152] Tillner-Roth, R.; Baehr, H.D.: An international standard formulation of the thermo-
dynamic properties of 1,1,1,2-tetraflouroethane (HFC-134a) covering temperatures
from 170 K to 455 K at pressure up to 70 MPa. J. Phys. Chem. Ref. data., Vol. 23,
1994, 657
[153] Tillner-Roth, R.; Li, J.; Yokozenki, A.; Sato, H.; Watanabe, K.: Thermodynamic
properties of pure and blended Hydroflourocarbons (HFCs) Refrigerants. Japan
Society of Refrigeration and Air Conditioning Engineers, 1997
[154] Urso, M.E.; Wadekar, V.V; Hewitt, G.F.: Heat transfer at the dryout and near
dryout regions in flow boiling. 12. Proc. of the 12. Int. Heat Tranasfer Conference,
Grenoble, Vol.3. 2002, 701-706.
[155] van Wijk, W.R.; Vos, A.S.; van Stralen, S.J.D.: Heat transfer to boiling binary
liquid mixtures. Chem. Engng. Sci., Vol. 5, 1956, 68-80
[156] Verma, H.K.; Sharma, C.P.; Mishra, M.P.: Heat transfer coefficients during forced
convective evaporation of R12 and R22 mixtures in annular flow regime. Proc. XV
Int. Congr. Refrig., Vol. II, 1979, 479-484
[157] VDI, VDI-GVC: VDI- Wärmeatlas. 8. Aufl., Springer-Verlag, Berlin, 1997

130 Bibliography
[158] Viera, G. A., L. L. Simpson and B. C. Ream 1993. Lessons Learned from the Ethy-
lene Oxide Explosion at Seadrift, Texas, Chemical Engineering Progress, August
1993.
[159] Voskresenskji, K.D.: Heat transfer in film condensation with temperature dependent
properties of the condensate. Izv. Akad. Nauk. USSR, 1948, 1023-1028
[160] Wadekar, V.V.: Convective heat transfer to binary mixtures in annular two-phase
flow. Proc. of the 10th Int. Heat Transfer Conference, Brington, Vol. 7, 1994, 557-562
[161] Wadekar, V.V.: Boiling hot issues-some resolved and some not-yet-resolved. Trans.
IchemE., Vol. 76, Part A, 1998, 133-142
[162] Weisman, J.; Duncan, D.; Bibson, J.; Crawford, T.: Effect of fluid properties and
pipe diameter on two phase flow pattern in horizontal pipelines. Int. J. Multiphase
flow, Vol. 5, 1979, 437-462
[163] Wettermann, M.: Wärmeübergang beim Sieden von Gemischen bei Zwangskonvek-
tion im horizontalen Verdampferrhor. Fortschritt-Berichte VDI- Reihe 3, Nr. 625,
VDI-Verlag GmbH, Düsseldorf, 1999
[164] Wongwises, S.; Disawas, S.; Kaewon, J; Onuari, C.: Two-phase evaporative heat
transfer coefficients of refrigerant HFC-134a under forced flow conditions in a small
horizotal tube. Int. Comm. Heat Mass Transfer, Vol. 27, No. 1, 2000, 35-48
[165] Yan, Y.; Lin, T.: Evaporation heat transfer and pressure drop of refrigerant R134a
in small pipe. Int. J. of Heat and Mass Transfer. Vol. 41, 1997, 4183-4194
[166] Younglove, B.A.; Ely, J.F.: Thermophysical properties of fluids: Methane, Ethane,
Propane, Isobutane and Normal Butane. J. of Physical Chemical Reference data,
Vol. 16, No. 4, 1987, 577-798
[167] Zahn, W.R.: Flow conditions when evaporating refrigerant R22 in air conditioning
coils. ASHRAE trans., Vol. 72, 1965, 82-89
[168] Zhang, L.; Hihara, E.; Saito, T.; Oh, J.-T.: Boiling heat transfer of a ternary refrig-
erant mixture inside a horizontal smooth tube. Int. J. of Heat and Mass Treansfer,
Vol. 40, No. 9, 1997, 2009-2017
[169] Zürcher, O.; Thome, J.R.; Favrat, D.: In tube flow boiling of R407C and R407C/oil
mixtures. Part I: microfin tube. ASHRAE Trans., Vol. 4, No. 4, 1998, 347-372

131
A Heat transfer coefficient
A.1 Single phase
A.1.1 Inside tube: Turbulent flow
Nu = CRea
Prb µ
µw
c
, (A.1)
where
Nu = hde
k
Nusselt number
Pr = Cpµ
k
Prandtl number
Re ρud
µ
Reynolds number
de
4A
P
hydraulic diameter
A cross-sectional area
P wetted perimeter
u fluid velocity
µw fluid viscosity at the tube wall temperature
k fluid thermal conductivity
Cp fluid specific heat
C =



0.021 gases
0.023 non-viscous liquid
0.027 viscous liquid
a = 0.8
b = 0.3 for cooling
b = 0.4 for heating
c = 0.14
A.1.2 Inside tube: Laminar flow
Nu = 1.86 RePr
d
L
1/3
µ
µw
0.14
, (A.2)
A.1.3 Shell side
For the shell side heat transfer coefficient there are a number of methods the include:
• Kern’s method
• Donohue’s method
• Bell-Delaware method
• Tinker’s method
Besides these methods there is some proprietary methods putout by various organization
for use by their member companies. A number of these method are based on one of the
above methods. Some are based upon a judicious combination of methods 3 and 4 above
and supplemented by further research data. Among the most popular of the proprietary
methods, judged by their large clientele are

132 A Heat transfer coefficient
• Heat Transfer Research Inc. (HTRI), Alliambra, california. This method is also
known as stream analysis method.
• Heat Transfer and Fluid Flow Service (HTFS), Engineering Science Division, AERE,
Harwell, United Kingdom Method.
In this work only Kern’s method is given below. Bell-Delaware method may be found in
Coulson and Richardson’s
Nu = 0.36Re0.55
Pr1/3 µ
µw
0.14
, (A.3)
where
Nu = hde
k
Nusselt number
Pr = Cpµ
k
Prandtl number
Re = Gde
µ
Reynolds number
de = 4A
P
hydraulic diameter
G = M
As
Mass flux
As = (pt−do)DslB
pt
fluid viscosity at the tube wall temperature
pt = pitch diameter
Ds = shell diameter
lB = Baffle spacing
Hydraulic diameter (Fig. A.1)
de =



p2
t −πd2
o/4
πdo
for square pitch
0.87p2
t /2−πd2
o/8
πdo/2
for equilateral triangular pitch
pt
do
Square pitch
pt
Equilateral triangular pitch
As
Cross-flow area
Figure A.1. Tube arrangement

A.2 Condensation 133
A.1.4 Plate heat exchanger
Nu = 0.26Re0.65
Pr0.4 µ
µw
0.14
, (A.4)
where
Nu = hde
k
= Nusselt number
Pr = Cpµ
k
= Prandtl number
Re = ρupde
µ
= Gde
µ
= Reynolds number
de = hydraulic diameter, taken as twice the gap between the plates
G = M
Af
= Mass flux
Af = cross-sectional area for flow
up = channel velocity
M = mass flow rate
A.2 Condensation
A.2.1 Condensation on vertical plate or outside vertical tube
hm = 0.943
k3
ρ∆ρgλ
µ∆TL
1/4
, (A.5)
where
hm = mean heat transfer coefficient
L = lenth of the plate or the vetical tube
k thermal conductivity of the saturated liquid film
ρ = liquid density
µ = liquid viscosity
λ = latent heat of evaporization
∆T = Ts − Tw temperature difference across the condensate film
g = acceleration due to gravity
Ts = saturation temperature of the condensate film
Tw = wall temperature
A.2.2 Condensation on external horizontal tube
hm = 0.725
k3
ρ∆ρgλ
µ∆Tdo
1/4
, (A.6)
where
do = out side diamter of the tube
A.2.3 Condensation on banks of horizontal tube
hm = 0.725
k3
ρ∆ρgλ
µ∆TJdo
1/4
, (A.7)

where
J = number of tubes in a row (Fig. ??)
In the above equation the condensate film properties save the latent heat of vaporization
are evaluated at the film temperature.
Tf =
Ts + Tw
2
, (A.8)
the latent heat of vaporization is evaluated at the condensate temperature. For the case
of subcooling or superheating the heat transfer coefficient is corrected by substituting the
corrected latent heat the heat transfer equation (Rohsenow et al. [121] and Carey [18]) in
Nusselt [109])
λ∗
= λ + 0.68cp∆T . (A.9)
A.2.4 Condensation inside horizontal tube
hm = 0.555
k3
ρ∆ρgλ
µ∆Td
1/4
, (A.10)
A.3 Two phase flow: Pure fluid
A.3.1 Steiner [140] correlation
Steiner [140] has considered the two phase heat transfer coefficient h as a combination of
the convective and the nucleate part using an asymptotic model as:
h = h3
n + h3
c
1/3
, (A.11)
where hn and hc is the nucleate and convective boiling heat transfer coefficient respectively.
The convective boiling heat transfer coefficient for a completely wetted tube (i.e. all types
of flow patterns save stratified and stratified-wavy flow) is calculated as
hc
hL0
=




(1 − ˙x) + 1.2 ˙x0.4
(1 − ˙x)0.01 ρL
ρG
0.37

 +


hG0
hL0
˙x0.01

1 + 8(1 − ˙x)0.7 ρL
ρG
0.67




−2



−0.5
. (A.12)
The heat transfer coefficients hL0 and hG0 are those of single phase flow, assuming that
the total mass velocity is pure liquid or pure vapor respectively. They are calculated in
the case of a fully developed turbulent flow from the Gnielinski [46] model
Nu =
(ξ/8)(Re − 1000)Pr
1 + 12.7(ξ/8)0.5(Pr2/3 − 1)
, (A.13)
taken in to account the respective dimensionless group NuL0, NuG0, ReL0, ReG0, Prl and
Prg. These dimensionless groups are defined as
NuL0/G0 =
hL0/G0d
kL/G
, (A.14)

A.3 Two phase flow: Pure fluid 135
ReL0/G0 =
˙md
µL/G
, (A.15)
PrL/G =
µL/Gcp,L/G
kL/G
. (A.16)
The friction factor is
ξ = (1.82logRe − 1.62)−2
. (A.17)
For a partial wetting of the tube (stratified or stratified-wavy flow) the average heat
transfer coefficient at the tube circumference under the thermal boundary condition of a
constant wall temperature is given as
hc = hwet(1 − Φ) + hGΦ , (A.18)
where hwet is the convective boiling heat transfer coefficient at the wetted part of the
tube and it is calculated by using equation A.12. In the non-wetted part of the tube,
the convective heat transfer coefficient hg is calculated from the Gnielinski [46] model
(equation A.13). In this case Re and Nu are defined with the hydraulic diameter of the
vapor-occupied part of the tube cross-section
dh = d
ϕ − sin ϕ
d + 2 sin(ϕ/2)
, (A.19)
where ϕ is the stratified angle. The Reynolds number is given as
ReG =
˙m ˙xdhyd
µG
, (A.20)
and
hG =
NuGkG
dhyd
. (A.21)
The void fraction is calculated using the Rauhani [117] model given as
ε =
˙x
ρG
(1 + 0.12(1 − ˙x))
˙x
ρG
+
1 − ˙x
ρL
+
1.18(1 − ˙x)[gσ(ρL − ρG)]1/4
˙mρ .
1/2
L
−1
(A.22)
The wetting boundary can be estimated (see Fig. A.2) from the void fraction as
ε =
fG
fG + fL
. (A.23)
With some mathematical manipulation of equation A.23 the non-wetted perimeter can
calculated iteratively from the following relationship
ϕ = 2πε + sinϕ , (A.24)
with the assumption that no bubbles in the liquid phase and no entrainment (hold-up) in
the vapor phase, the scaling parameter Φ of equation A.18 can thus be calculated as
Φ =
ϕG
2π
, (A.25)
where ϕG = 0.5ϕ.

ϕ
d
hfL
fG
Ui
UG
UL
Figure A.2. Cross-section and perimeter parts of the vapor flow in a horizontal tube.
The local nucleate boiling heat transfer coefficient hnb of a horizontal tube is estimated
as
hnb
ho
= ψCf
˙q
˙qo
n(pr)
F(pr)F(Ra)F(d)F( ˙m, ˙x) . (A.26)
The value with a subscript ”o” is a reference value.
The pressure function is given as
F(pr) = 2.692p0.43
r +
1.6p6.5
r
1 − p4.4
r
, (A.27)
and the mass flux function is given as
F( ˙m, ˙x) =
˙m
˙mo
0.25

1 − p0.1
r
˙q
qcr,nb
0.3
˙x

 , (A.28)
where
˙qcr,cb = 2.79 ˙qcr,0,1p0.4
r (1 − pr) . (A.29)
The critical value of ˙qcr,0,1 at a reduced pressure pr of 0.1 is given as
˙qcr,0.1 = 0.13∆hV,0ρ0.5
G,0[σog(ρL,0 − ρG,0)]0.25
. (A.30)
The function for the effect of surface roughness and tube diameter is F(Ra) =(Ra/Rao)0.133
and F(d)=(do/d)0.5
respectively. The pressure dependence of the heat flux exponent n(pr)
can be predicted as
n(pr) = 0.9 − 0.3p0.3
r . (A.31)
The experimental value of the specific constant Cf for a number of substances is be found
in VDI-Wärmeatlas[157], for example for water it is 0.72. In absence of an experimental
value it can be estimated as
Cf = 0.789
M
MH2
0.11
, (A.32)
where M is the molecular weight and MH2 = 2.016. The correction factor ψ for a stratified
and a stratified-wavy flow pattern under the thermal boundary condition of a constant

wall temperature is 0.86 for all other type of flow patterns it is taken as unity (VDI-
Wärmeatlas[157]).
Table A.1 shows the reference factors for the nucleate boiling heat transfer coefficient for
R134a and R290.
Table A.1. Values of the reference parameters used in evaluation of the local nucleate boiling
heat transfer coefficient.
Refrigerant ho ˙qo Rao do
W/m2
K W/m2
m m
R134a 3,500 20,000 10−6
0.01
R290 4,000 20,000 10−6
0.01
A.3.2 Kattan et al. [77] correlation
For a stratified-wavy flow pattern or annular flow pattern with a partial dryout the two
phase heat transfer coefficient is
h =
ϕdryhG + (2π − ϕdry)hwet
2π
. (A.33)
The vapor heat transfer coefficient hG is determined by using the Dittus-Boelter [33]
correlation as
hG = 0.023Re0.8
G P0.4
rG
kG
d
, (A.34)
with Reynold number given as
ReG =
˙m ˙xd
εµG
, (A.35)
where ε is the void fraction given by the Rauhani [117] model (equation A.22). The heat
transfer coefficient on the wetted portion of the tube is
hwet = 3
h3
n + h3
c . (A.36)
The nucleate boiling heat transfer coefficient hn is given by the Cooper [27] model as
hn = 55p0.12
r (−0.4343 ln pr)−0.55
M−05
˙q . (A.37)
The convective heat transfer coefficient is given by a modified form of the Dittus-Boelter
[33] model as
hc = 0.0133Re0.69
L P0.4
rL
kL
d
. (A.38)
The liquid Reynolds number is given as
ReL =
4 ˙m(1 − ˙x)δ
(1 − ε)µG
. (A.39)
where δ is the liquid film thickness it is given as
δ =
πd(1 − ε)
2(2π − ϕdry)
, (A.40)

where ϕdry is
ϕdry = ϕstrat
( ˙mwavy − ˙m)
( ˙mwavy − ˙mstrat)
, (A.41)
where ϕstrat is calculated iteratively from equation A.24. The mass flux under a stratified
and wavy flow pattern is
˙mstrat =
(226.3)2
fLf2
GρG(ρL − ρG)µLg cos Θ
0.3164(1 − ˙x)1.75π2µ0.25
L
, (A.42)
and
˙mwavy =
16f3
GgdρLρG
˙x2π2(1 − (2hL − 1)2)0.5
π2
25h2
L
(1 − x)F1( ˙q)
×
We
Fr
F2( ˙q)
L
+
1
cos Θ
0.5
+ 50 ,
(A.43)
respectively. The parameters fL, fG, hL are defined in Fig.A.2. Θ is the angle of inclination
to the horizontal and
F1( ˙q) = 646.0
˙q
˙qcrit
2
+ 64.8
˙q
˙qcrit
; F2( ˙q) = 18.8
˙q
˙qcrit
+ 1.023 . (A.44)
The stratified-wavy flow model is also valid for the stratified flow patten with ϕstrat
replacing ϕdry and for the annular flow condition with ϕdry is set to zero and the film
thickness δ is set to (1 − ε)d/4.
A.3.3 Kandlikar [70] correlation
The flow boiling heat transfer coefficient for a pure fluid is given by Kandlikar [70] as
h = max(hn, hc) , (A.45)
wher the subscript n and c in equation A.45 refers to the nucleate and convective boiling
respectively. The convective and the nucleate boiling part is given as
hn = 0.6683Co−0.2
(1 − ˙x)0.8
hL0f(FrL0) + 1058.0Bo0.7
(1 − ˙x)0.8
FFlhL0 , (A.46)
and
hc = 1.136Co−0.9
(1 − ˙x)0.8
hL0f(FrL0) + 667.2Bo0.7
(1 − ˙x)0.8
FFlhL0 , (A.47)
respectively, where FrL0 is the liquid Froude number, Bo is the boiling number and Co
is the convection number. These dimensionless groups are defined as
FrL0 =
˙m
ρLgd
, Bo =
˙q
˙m∆hv
, Co =
ρG
ρL
0.5
1 − ˙x
˙x
0.8
. (A.48)
The function f(FrL0) is defined as
f(FrL0) = (25FrL0)0.324
FrL0 < 0.04 ,
f(FrL0) = 1 FrL0 ≥ 0.04 ,
where FFl is a fluid-surface parameter related to the nucleation characteristic. For all
type of fluids flowing in a stainless tube it is taken as 1. The single phase heat transfer

coefficient hL0 is obtained from the Petukhov and Popov [114] correlation or Gnielinski
[46] correlation. The Petukhov and Popov [114] correlation is valid in the range of 0.5 ≤
PrL ≤ 2000 and 104
≤ ReL0 ≤ 5 × 106
and it is given as
NuL0 =
hL0d
k
=
ReL0PrL(ξ/2)
1.07 + 12.7(P
2/3
rL − 1)(ξ/2)0.5
. (A.49)
The Gnielinski [46] correlation (equation A.13) is valid in the range of 0.5 ≤ PrL ≤
2000 and 2300 ≤ ReL0 ≤ 5 × 104
. The friction factor ξ in equation A.49 is given by
equation A.17.
A.3.4 Chen [19] correlation
Chen [19] postulated that the heat transfer coefficient is made of two parts: a) a micro-
convective (or nucleate boiling) portion hn and b) a macro-convective (or forced convec-
tive) portion hc as
h = hcF + hnS , (A.50)
where hc is calculated using the Dittus and Boelter [33] correlation as
hc = 0.023
kL
d
Re0.8
L Pr0.4
L , (A.51)
where
ReL =
(1 − ˙x) ˙md
µL
, PrL =
cpLµL
kL
, (A.52)
The suppression factor for the convection part is
F =



1 if 1/Xtt > 0.1
2.35 1
Xtt
+ 0.213
0.736
if 1/Xtt ≤ 0.1
,
and the Martinelli parameter Xtt is given as
X =
1 − ˙x
˙x
0.875
ρG
ρL
0.5
µL
µG
0.125
. (A.53)
The nucleate boiling heat transfer coefficient is
hn = 0.00122
k0.79
L c0.45
p,L ρ0.49
L
σ0.5µ0.29
L ρ0.24
G ∆h0.24
V
∆T0.24
sat ∆p0.75
sat , (A.54)
where
∆Tsat = Tw − Ts; ∆psat = p(Tw) − p(Ts); Retp = ReLF1.25
. (A.55)
The suppression factor for the nucleate part is
S =
1
1 + 2.53 × 10−6Retp
. (A.56)

A.3.5 Gungor and Winterton [52] correlation
The Gungor and Winterton [52] correlation is a modified form of the Chen [19] correlation
given by equation A.50 with the nucleate boiling calculated from the Cooper [27] corre-
lation given by equation A.37. The suppression factor for the convection part is defined
as
F =



(1 + 24, 000Bo1.16
+ 1.37(1/xtt)0.86
)Fr
(0.1−2FrL)
L if Fr < 0.05
1 + 24, 000Bo1.16
+ 1.37(1/xtt)0.86
if Fr ≥ 0.05
,
and the suppression factor for the nucleate part is
S =



(1 + 0.00000115F2
ReL)−1
Fr
1/2
L if Fr < 0.05
(1 + 0.00000115F2
ReL)−1
if Fr ≥ 0.05
,
The convective boiling part is calculated from the Dittus-Boelter [33] correlation (equation
A.51).
A.3.6 Shah [130] correlation
The Shah [130] correlation is given as
h = max(hc, hn) , (A.57)
where the subscript n and c in equation A.57 refers to the nucleate and convective boiling
respectively. The convective heat transfer coefficient is defined as
hc = 1.8hLN−0.8
, (A.58)
where
N =



Co FrL > 0.04
0.38Fr−0.4
L Co FrL < 0.04
,
where hL is calculated using the Dittus-Boelter [33] correlation (equation A.51). The
nucleate boiling heat transfer coefficient is calculated as follows
• For N > 1
hn =



230hLBo0.5
Bo > 0.0003
1 + 46hLBo0.5
Bo < 0.0003
.
• For 1 > N > 0.1
hn = FhLBo0.5
exp(2.74N−0.1
) . (A.59)
• For N < 0.1
hn = FhLBo0.5
exp(2.47N−0.15
) , (A.60)
where
F =



14.7 Bo > 0.0011
15.43 Bo < 0.0011
.

A.3.7 Schrock and Grossman [129] correlation
A very simple correlation is given by Schrock and Grossman [129] as
h = 1.91hL 104
× Bo + 1.5
1
Xtt
2/3 0.6
, (A.61)
where hL is calculated using Dittus-Boelter [33] correlation equation A.51.
A.3.8 Dembi et al. [30] correlation
The Dembi et al. [30] correlation is based on the asymptotic model given by equation
A.11 with the nucleate and convection part given as
hn = 23388.5
kL
d
˙q
ρG∆hV
0.64
gd
∆hV
0.27
˙m2
d
ρL∆hV
0.14
, (A.62)
and
hc = 0.115
kL
d
˙x4
(1 − ˙x)2 0.11 ˙m2
∆hV
ρLgσ
0.14
P0.27
rL , (A.63)
respectively. The parameter is deﬁned as
= 0.36 × 10−3
p−1.4
r . (A.64)
A.3.9 Klimenko [84] correlation
The Klimenko [84] correlation is based on the asymptotic model given by equation A.11
with the convection part given by the Dittus-Boelter [33] correlation equation A.51 and
the nucleate boiling is
hn =
hn1 NCB < 1.6 × 104
hn2 NCB > 1.6 × 104 ,
where
hn1 = 7.4 × 10−3 kw
kL
0.15
Pe0.6
K0.5
p Pr
−1/3
L , (A.65)
hn2 = 0.087
kL
b
kw
kL
0.09
Re0.6
m
ρG
ρL
0.2
Pr
1/6
L , (A.66)
Pe =
qb
∆hV ρGaL
, Kp =
p
σg(ρL − ρG)
, b =
2σ
g(ρL − ρG)
, (A.67)
Rem =
wmb
νL
, wm =
˙m
ρL
1 + x
ρL
ρG
− 1 , Re∗ =
qb
∆hV ρGνL
, NCB =
Rem
Re∗
ρL
ρG
.
(A.68)

A.3.10 Jung et al. [64] correlation
The Jung et al. [64] correlation is a modified form of the Chen [19] correlation. The
convection heat transfer coefficient is calculated using the Dittus-Boelter [33] correlation
(equation A.51) and the nucleate part is calculated from the Stephan and Abdelsalm in
VDI-Wärmeatlas [157] correlation as
hn = 207
kL
b.d
˙q(b.d)
kLTs
0.745
ρG
ρL
0.581
P0.533
rL , (A.69)
where
(b.d) = 0.511
2σ
g(ρL − ρG)
0.5
, (A.70)
F = 2.37 0.29 +
1
Xtt
, (A.71)
S =



4048X1.22
tt Bo1.13
Xtt < 1
2.0 − 0.1X−0.28
tt Bo−0.33
1 ≤ Xtt ≤ 5
.
A.4 Two phase flow: Mixture
A.4.1 Steiner [140] correlation
Steiner [140] has extended his pure component asymptotic model to mixture. The nucleate
part of the heat transfer coefficient is suppressed using the Schlünder [126] suppression
factor for the nucleate boiling. The Schlünder [126] suppression factor is based on the
heat and mass transfer laws it is defined as
Fn = 1 +
hid,n
˙q
(Tb,k − Tb,j)(yj − xj) 1 − exp
Boq
ρL∆hV βL
, (A.72)
where Tb is the saturated (boiling) temperature of the pure component, the index j and
k stands for the more volatile and less volatile component respectively. βL/B0 = 5 × 105
is the mass transfer coefficient. The ideal nucleate boiling heat transfer coefficient for a
mixture hid,n is calculated from the heat transfer coefficient of pure components as
hid,n =
xi
hi,n
−1
, (A.73)
and Bo/βL = 5.103
and ρL and ∆hV is the ideal density and enthalpy of evaporation of
the mixture respectively. x and y is the liquid and vapor mole fraction of the more volatile
component respectively.
The same approach applies also to the convective part for the liquid-liquid immiscible
mixture. That is to say for a liquid-liquid miscible mixture the convective suppression
factor made analogous to that for the nucleate boiling heat transfer coefficient as
Fc = 1 +
hid,c
˙q
(Tb,k − Tb,j)(yj − xj) 1 − exp
Boq
ρL∆hV βL
. (A.74)

A.4 Two phase flow: Mixture 143
A.4.2 Kandlikar [71] correlation
Kandlikar [71] has extended his pure component correlation (Kandlikar [70]) to mixtures
as
• Region I: Near-azeotropic region
h = max(hn, hc) , (A.75)
where hn and hc is obtained from equation A.77 and equation A.47 respectively
using the mixture properties.
• Region II: Moderate diffusion-induced suppression region
h = hc , (A.76)
where hc is given by equation A.77 with the properties of the mixture.
• Region III: Severe diffusion-induced suppression region: 0.03< V1 < 0.2 and Bo ≤
1E−4
; V1 ≥ 0.2
h = 1.136Co−0.9
(1 − ˙x)0.8
hL0
f(FrL0) + 667.2Bo0.7
(1 − ˙x)0.8
FFlhL0FD , (A.77)
where
V1 =
cpL
∆hV
a
D12
0.5
|y − x|
dT
dx
, (A.78)
FD =
0.678
1 + V1
. (A.79)
A.4.3 Bennett and Chen [8] correlation
Bennett and Chen [8] has extended the Chen [19] correlation (equation A.50) for mix-
ture. Here both the convective and the nucleate parts are suppressed. The convection
part which is calculated for the original Chen [19] correlation with mixture properties is
suppressed using the following suppression factor
Fc =
Tw − Tph
Tw − Ts
, (A.80)
where Tw, Tph, and Ts is the wall, equilibrium temperature and saturation temperature
respectively. The nucleate part is also calculated using the original Chen [19] model for
the pure substance with mixture properties. It suppressed using the the suppression factor
given by equation A.79.
A.4.4 Palen [111] correlation
Palen [111] has extended the original Chen [19] correlation for pure component (equation
A.50) to mixture similar to the Bennett and Chen [8] correlation. However, only the
nucleate part is suppressed using the following suppression factor
Fd = exp(−0.027∆Tbp) , (A.81)
where ∆Tbp is difference between the dew and bubble point temperature of the mixture.

A.4.5 Jung et al. [64] correlation
Jung et al. [64] have extended their pure substance correlation to the mixture. The nu-
cleate boiling heat transfer coeﬃcient is replaced by the ideal one given by equation A.73.
The convective part is suppressed using the following suppression factor
Fc = 1.0 − 0 − 35|y1 − x1|1.56
. (A.82)
For the nucleate part the following suppression factor is employed
Fn =
1
{[1 + (b2 + b3)(1 + b4)](1 + b5)}2 , (A.83)
where
b2 = (1 − x1) ln
1.01 − x1
1.01 − y1
+ x1 ln
x1
y1
+ |y1 − x1|1.5
, (A.84)
b3 =



0 x1 ≥ 0.01
x1
y1
0.1
− 1 x1 < 0.01
,
b4 = 152
p
pc,1
0.66
, (A.85)
b5 = 0.92|y1 − x1|0.001 p
pc,1
0.66
, (A.86)
and
x1
y1
= 1 for x1 = y1 = 0 ,
x1 and y1 is the liquid and vapor mole fraction of the more volatile component respec-
tively. p and pc,1 is system pressure and critical pressure of the more volatile component
respectively.

145
B Pressure drop
B.1 Single phase
various approaches have been taken, for example the frictional pressure gradient is given
as
−
dp
dz f
=
4τo
d
=
4f ˙m2
2dρ
, (B.1)
where ˙m is the mass flux in kg/m2
s and f is the friction factor calculated using a Blasius-
type model as
f =



0.3164
Re0.25 Re ≥ 2320
64
Re
Re < 2320 .
Integration of equation B.1 yields
∆p =
4f ˙m2
2ρ
L
d
, (B.2)
B.2 Two phase
In flow boiling, the temperature drops in the direction of flow as a result of the pressure
drop. This results in a change in the driving force (temperature difference) for the heat
transfer along the flow path. Thus beside the heat transfer coefficient, knowledge of the
pressure drop is of paramount importance in the design of the evaporator. In the present
work the pressure drop is measured simultaneously with the heat transfer coefficient along
the test section.
The momentum balance implies that the two phase pressure gradient is composed of three
components as
dp
dz
=
dp
dz f
+
dp
dz a
+
dp
dz h
, (B.3)
where dp/dz, (dp/dz)f , (dp/dz)a and (dp/dz)h is the total, friction, acceleration and
hydrostatic pressure gradient respectively. For a horizontal tube the hydrostatic pressure
gradient diminishes. The acceleration pressure drop is caused by the change in momentum
in both the liquid and vapor phases. The change in the momentum stems from the change
in the velocity of the two phases, which is brought about by the added (or withdrawn)
heat to/from the test section. For the case of adiabatic flow the acceleration pressure drop
diminishes for ∆pa/ps → 0 (Baehr and Stephan [3]), where ps is the saturation pressure.
There exist in the literature a number of approaches for modelling the change in the static
pressure drop due to acceleration. The most widely accepted models include homogenous
or separated flow models. The separated flow model is also widely known as the het-
erogenous model. In the homogenous model the static pressure drop due to acceleration
is
−
dp
dz a
= ˙m2 d
dz
˙x
1
ρL
−
1
ρG
+
1
ρL
. (B.4)
The energy balance in a small unit length dz along the test tube yields
d ˙x
dz
=
4 ˙q
˙m∆hvd
. (B.5)

146 B Pressure drop
Substitution of equation B.5 into equation B.4 yields the pressure drop due to acceleration
as
∆pa =
4 ˙q ˙m
d∆hvρG
1 −
ρG
ρL
∆L . (B.6)
In the separated flow model the static pressure drop due to acceleration can be derived
from the momentum balance as
−
dp
dz a
= ˙m2 d
dz
˙x2
ερG
+
(1 − ˙x)2
(1 − ε)ρL
. (B.7)
Integration of equation B.7 between the inlet i and outlet o of the test section yields
−∆pa = −(po − pi)a = ˙m2 ˙x2
2
εoρG,o
+
(1 − ˙xo)2
(1 − εo)ρL,o
−
˙x2
i
εiρG,i
−
(1 − ˙xi)2
(1 − εi)ρL,i
. (B.8)
The void fraction ε may be obtained using the Rauhani [117] model which is given as:
ε =
˙x
ρG
(1 + 0.12(1 − ˙x))
˙x
ρG
+
1 − ˙x
ρL
+
1.18(1 − ˙x)[gσ(ρL − ρG)]1/4
˙mρ
1/2
L
−1
, (B.9)
where ρL and ρG is the liquid and vapor density respectively, which are calculated from the
fundamental equation of state of Tillner-Roth and Baehr [152] for R134a. g is acceleration
due to gravity, σ is the surface tension, ˙m is the mass flux and ˙x is the quality. The surface
tension is calculated using the method of Lucus [92] given in VDI-Wärmeatlas [157].
various approaches have been taken, for example in homogenous or separated flow models.
In the homogenous model the frictional pressure gradient is given as
−
dp
dz f
=
4τo
d
=
2ξ ˙m2
dρH
, (B.10)
where ξ is the two phase friction factor calculated by a Blasius-type model as
ξ =



0.3164
Re0.25 Re ≥ 2320
64
Re
Re < 2320 .
and the homogenous densityρH is given as
1
ρH
=
1 − ˙x
ρL
+
˙x
ρG
. (B.11)
The two phase Reynolds number Re is
Re =
˙md
ηTP
, (B.12)
where ηTP is a two-phase viscosity. A variety of methods have been proposed to calculate
the two phase viscosity, a commonly used one being that proposed by McAdams et al. [95]
1
ηTP
=
1 − ˙x
ηL
+
˙x
ηG
, (B.13)

B.2 Two phase 147
where ηL and ηG are the liquid and vapor viscosity.
In the separated flow model the two phase frictional pressure drop is related to that for
single phase as
dp
dz f
=
dP
dz f,L/G
ΨG/L , (B.14)
where Ψ is the two phase multiplier. There exist a number of correlations for the prediction
of Ψ. These include Friedel [42], Chishlom [22] and Lockhart and the Martinelli [91] model.
These models are presented in Appendix B. There exists a number of correlations for the
prediction of the two phase multiplier Ψ of the separated flow model. These models are
presented in the following subsections.
B.2.1 Friedel [42] model
ΨL0 = E +
3.24FH
Fr0.045We0.035
, (B.15)
where
E = (1 − ˙x)2
+ ˙x2 ρLfG0
ρGfL0
, (B.16)
F = ˙x0.78
(1 − ˙x)0.24
, (B.17)
H =
ρL
ρG
0.91
µG
µL
0.19
1 −
µG
µL
0.7
, (B.18)
Fr =
˙m2
gdρ2
H
, (B.19)
We =
˙m2
d
σρH
, (B.20)
d is tube diameter, σ is the surface tension and H is the homogenous density given by
equation B.11. fG0 and fL0 are the friction factors defined by a Blasius-type model as
fL0/G0 =
0.079
Re0.25
L0/G0
, (B.21)
where Re = ˙md/µ. The range of the validity of the Friedel [42] model is µL/µG < 1000
B.2.2 Lockhart and Martinelli [91] model
In the Lockhart and Martinelli [91] model the two phase friction multiplier is
ψ2
L = 1 +
C
X
+
1
X2
, (B.22)
ψ2
G = 1 + C.X + X2
, (B.23)
where X is the Martinelli parameter and the value of the coefficient C is given in Table B.1.
The range of the applicability of the Lockhart and Martinelli [91] correlation is µL/µG >1000
and ˙m <100 kg/m2
s.

148 B Pressure drop
Table B.1. Value of C for the Lockhart and Martinelli [91] correlation.
Liquid Gas Subscript C
Turbulent Turbulent tt 20
Viscous Turbulent vt 12
Turbulent Viscous tv 10
Viscous Viscous vv 05
B.2.3 Chisholm [22] model
In the Chisholm [22] model the two phase friction multiplier is
ΨL0 = 1 + (Y 2
− 1) B ˙x(2/n−1)
(1 − ˙x)(2/n−1)
+ ˙x1−n
, (B.24)
where
Y 2
=
(dpf /dz)G0
(dpf /dz)L0
, (B.25)
n is 0.25 for a Blasius model. The parameter B is given by
B =
55
˙m1/2
0 < Y < 9.5 , (B.26)
B =
520
Y ˙m1/2
9.5 < Y < 28 , (B.27)
B =
15000
Y 2 ˙m1/2
28 < Y . (B.28)
The range of the validity of the Chisholm [22] correlation is µL/µG > 1000 and ˙m > 100
kg/m2
s.

149
C Physical properties
The fluid physical properties required for heat exchanger design are divided in thermo-
dynamic and trasport properties. The transport properties include viscosity, thermal
conductivity, surface tension and diffusion coefficient are generally calculated from the
existing correlations (Pery and Coulson). The thermodynamic properties include dem-
sity, specific heat temperature, pressure (vapor), enthalpy, latent heat of evaporation.
Beside the fluid properties the thermal conductivity of the material is necessary for the
evaluation of heat transfer coefficient. The thermodynamic properties are evaluated using
critical tables.
C.1 Physical properties: Pure fluid
C.1.1 Specific heat
The specific heat of the ideal gas is given in as
Cp = CPV APA + (CPV APB)T + (CPV APC)T2
+ (CPV APD)T3
(C.1)
Where T is in K and CPVAPA, CPVAPB, CPVAPC, CPVAPD are constant in ideal
gas heat capacity. These constant are given in Appendix A for organic and inorganic
compounds.
C.1.2 Vapor pressure
The vapor pressure is generally predicted using Antonie equation as
ln p = ANTA −
ANTB
T + ANTC
(C.2)
where T is in K and ANTA, ANTB,ANTC are Anonie equation constant. These constant
are given in Appendix D for organic and inorganic compounds.
C.1.3 Liquid viscosity
The liquid viscosity is given as:
log µ = V ISA
1
T
−
1
V ISB
(C.3)
where VISA, VISB are constants in the liquid viscosity equation. These constant are
given in Appendix D for organic and inorganic compounds.
C.1.4 Vapor dynamic viscosity VDI-Wärmeatlas [157]
Lucas and Luckas [92] in VDI-Wärmeatlas [157] have recommended the following proce-
dure for the calculation of the vapor viscosity.
η = (ηξ)r
FpFQ
1
ξ
, (C.4)
for Tr ≤ 1 and pr ≤ ps/pc
(ηξ)r
= 0.600 + 0.760pα
r + (6.990pβ
r − 0.6)(1 − Tr) , (C.5)
with
α = 3.262 + 14.98p5.508
r and β = 1.390 + 5.746pr , (C.6)

150 C Physical properties
for 1≤ Tr ≤ 40 and 0≤ pr ≤ 100
(ηξ)r
= (ηo
ξ) 1 +
ApE
r
BpF
r + (1 + CpD
r )−1
, (C.7)
where ηo
is the low pressure viscosity given as
ηo
ξ = [0.807T0.618
r − 0.357 exp(−0.449Tr) + 0.340 exp(−4.058Tr) + 0.018]Fo
p Fo
Q , (C.8)
and ξ is given as
ξ =
[Tc]1/6
[R]1/6
[Na]1/3
[M]1/2[pc]2/3
, (C.9)
where Na is the Avagadro number in kmol. The coefficients of equation C.7 are given as
A =
a1
Tr
exp(a2Tγ
r ) , (C.10)
B = A(b1Tr − b2) , (C.11)
C =
c1
Tr
exp(c2Tδ
r ) , (C.12)
D =
d1
Tr
exp(d2Tr ) , (C.13)
E = 1.3088 , (C.14)
F = f1 exp(f2Tς
r ) . (C.15)
The coefficients a, b, c, d, e, and f are given in Table C.1
Table C.1. Coefficients of the correlation used for the prediction of the vapor dynamic viscosity.
a1 1.245.10−3
a2 5.1726 c1 0.4489 c2 3.0578 γ -0.3286
b1 1.6553 b2 1.2723 d1 1.7368 d2 2.2310 δ -37.7332
f1 0.9425 f2 −0.1853 ς 0.4489 -7.6351
Fp = 1 + (Fo
p − 1)
(ηξ)r
ηoξ
−3
, (C.16)
and
FQ = 1 + (Fo
Q − 1)
(ηξ)r
ηoξ
−1
− 0.007 ln
(ηξ)r
ηoξ
4
, (C.17)
where Fo
p and Fo
Q is low-pressure polarity and quantum factors respectively. These factors
are
Fo
p = 1 , 0 ≤ µr < 0.022 , (C.18)
Fo
p = 1 + 30.55(0.292 − Zc)1.7
, 0.022 ≤ µr < 0.075 , (C.19)
Fo
p = 1 + 30.55(0.292 − Zc)1.7
(|0.96 + 0.1(Tr − 0.7)|) , 0.075 ≤ µr , (C.20)
where Zc is the critical compressibility factor and Fo
Q = 1.0 for all substances other than
He, H2 and D2 . The reduced dipole moment µr is given as
µr =
µ2
pc
(kTc)2
, (C.21)
where the dipole moment µ for the gases is given in VDI-Wärmeatlas [157]

C.1 Physical properties: Pure fluid 151
C.1.5 Dynamic viscosity of Fenghour et al. [40]
The functional form of the liquid and vapor viscosity of ammonia as given by Fenghour
et al. [40] is
η = ηo(T) + η1(T)ρ + η2(ρ, T) , (C.22)
The first term of the expansion is the dilute gas term which is given as
ηo(T) = 100
0.021357
0.29572
(MT)1/2
exp(Ω)
, (C.23)
where M is the molecular weight in g/mol, T is the temperature in K. The collision
integral Ω is defined as
Ω(T) = C(1) + C(2) log
kT
+
4
n=3
C(n) log
kT
n
, (C.24)
where /k=386 K and the value of the coefficient C is given in table C.2.
Table C.2. Coefficients for the Collision integral Ω (equation C.24).
C(1) 4.9931822 C(2) -0.61122364 C(3) 0.18535124 C(4) -0.1116094
The second term of equation C.22 represents the contribution of the moderately dense
fluid
η1(T) = Fv(T)ηo(T)ρ , (C.25)
where
Fv(T) = C



A(1) +
13
i=2
A(i) log
kT
−(i−1)
2



, (C.26)
where C=0.6022137/0.29573
and the value of the coefficient A is given in table C.3
Table C.3. Coefficients of equation C.26.
i A i A
1 -0.17999496×101
2 0.466692621×102
3 -0.53460794×103
4 0.33604074×104
5 -0.13019164×105
6 0.33414230×105
7 -0.58711743×105
8 0.71426686×105
9 -0.59834012×105
10 0.33652741×105
11 -0.12027350×105
12 0.24348205×104
13 -0.120807957×103
The third term in the viscosity equation C.22 is the contribution of the dense gas
η2(ρ, T) =
3
i=1
F(i, T)ρi+1
, (C.27)

where
F(i, T) =



1 0.219664285 kT
2
− 0.83651107 × 10−1
kT
4
2 0.17366936 × 10−2
− 0.83651107 × 10−2
kT
3 0.167668649 × 10−3
kT
2
− 0.149710093 × 10−3
kT
3
+
0.77012274 × 10−4
kT
4
The Fenghour et al. [40] correlation for the vapor viscosity of ammonia has an uncertainty
of 2% in the temperature range of T < Tc.
C.1.6 Surface tension
Lucas and Luckas [92] in VDI-Wärmeatlas [157] have recommended the following corre-
lation for the calculation of the surface tension
σ = p2/3
c T1/3
c
1 − Tr
a
m
b , (C.28)
where the reduced pressure and temperature are defined as
pr =
p
pc
, Tr =
T
Tc
, , (C.29)
respectively.
For a polar fluid like R134a the following quantities are valid
a = 1 , (C.30)
b = 0.1574 + 0.359ω − 1.769X − 13.69X2
− 0.510ω2
+ 1.298ωX , (C.31)
m = 1.210 + 0.5385ω − 14.61X − 32.07X2
− 1.656ω2
+ 22, 03ωX , (C.32)
X = lgpsr(Tr = 0.6) + 1.70ω + 1.552 . (C.33)
where ω is the acentric factor and it is given by Pitzer in VDI-Wärmeatlas [157] as The
surface tension given by equation C.28 is in 10−5
N/cm. Its level of uncertainty as given
by Reid et al. [118] is 1.2 % in the range of the reduced temperature of 0.56 ≤ Tr ≤ 0.63.
C.1.7 Thermal conductivity for liquids
k = 3.65 × 10−5
Cp
ρ
M
1/3
. (C.34)
where k thermal conductivity W/moC, M is the molecular mass, Cp speific heat capacity
(kJ/kg oC), ρ density (kg/m
3
)
C.1.8 Thermal conductivity for gases
k = µ Cp +
10.4
M
. (C.35)
where k thermal conductivity W/m o
C, M is the molecular mass, Cp specific heat capacity
(kJ/kg o
C), µ viscosity in (mNs/m2
)

C.2 Physical properties: Mixture 153
C.1.9 Specific enthalpy
For the vapor phase, the deviation of the specific enthalpy from the ideal state can be
illustrated using Redlich-Kwong equation written as
z3
+ z2
+ z(B2
+ B − A) = 0 . (C.36)
where z is the compressibilty factor defined as
z =
pv
RT
. (C.37)
and
A =
aP
R2T2.5
, B =
bp
RT
. (C.38)
h = ho + RT +
v
0
T
dP
R2T2.5dT
− p dv . (C.39)
C.2 Physical properties: Mixture
C.2.1 Liquid dynamic viscosity of mixtures
For a liquid mixture which contains one or more polar constituents Reid et al. [118]
recommended the following model for the calculation of the mixture liquid viscosity
ln ηm =
n
i=1
xi. ln ηL,i + 2.x1.x2.G12 , (C.40)
where xi is the mole fraction of the component i, ηL,i is the viscosity of the component i
in kg/ms and G12 is an adjustable parameter normally obtained from experimental data.
For a polar-nonpolar mixture G12= -0.22. The Reid et al. [118] model give the thermal
conductivity with a mean error of less then 5%.
C.2.2 Vapor dynamic viscosity of mixtures
The viscosity of a gas mixture can be approximated by using the principle of the kinetic
theory (Reid et al. [118]) as
ηm = ηo
m + ∆η , (C.41)
where ηo
m is the mixture gas viscosity at a low pressure and ∆η is a correction factor for
the high pressure viscosity
ηo
m =
n
i=1
yiηG,i
n
j=1 yiφij
, (C.42)
where yi is the mole fraction of the component i and ηi is the viscosity of the pure
component i. φij is a parameter which may be estimated as
φij =
1 + (ηG,i/ηG,j)0.5
(Mj/Mi)0.25
2
[8(1 + Mi/Mj)]0.5
, (C.43)
φji =
ηG,j
ηG,i
Mj
Mi
φij . (C.44)

The high pressure correction term is estimated as
∆η =
0.497.10−6
exp(1.439ρr,m) − exp(−1.111ρ1.858
r,m )
T
1/6
c,m M−0.5
m p
−2/3
c,m
. (C.45)
The pseudo critical properties of the mixture are calculated as
Tc,m =
j=1
yjTc,j, υc,m =
j
yjυc,j, Zc,j =
pc,jυc,j
RTc,j
, Zm =
j
yjZc,j, (C.46)
Mm =
j=1
yjMj, ρc,m =
Mm/1000
υc,m
, ρr,m =
ρm
υc,m
, pc,m =
RTc,mZc,m
υc,m
, (C.47)
where T is in K, p is in Mpa, υc,m is in m3
/kmol, ρr,m is in kg/m3
, M is in g/mol and ηm
is in kg/ms. The error associated with this model is seldom exceeded 3 to 4% (Perry and
Green [112]).
C.2.3 Liquid thermal conductivity of mixtures
Reid et al. [118] have recommended a Filippov-like model for the prediction of the thermal
conductivity of a liquid mixture as
λm =
2
i=1
XiλL,i − 0.72X1X2|λL,2 − λL,1| , (C.48)
where X1 and X2 is the weight fraction of the component 1 and 2 respectively and λ1 and
λ2 is the thermal conductivity of the component 1 and 2 in W/mK respectively.
C.2.4 Vapor thermal conductivity of mixtures
The thermal conductivity of a low-pressure gas mixture can be determined from the
relationship given by Reid et al. [118]
λG,m =
n
i=1
yiλG,i
n
j=1 yiAij
, (C.49)
where λG,m is the low-pressure gas mixture thermal conductivity, λG,i is the low-pressure
thermal conductivity of the pure component i. For a binary mixture of two non-polar
gases or a non-polar and a polar gas, Aij may be calculated by the model given by Perry
and Green [112] as
Aij =
1 + (λtr,i/λtr,j)0.5
(Mj/Mi)0.25
2
[8(1 + Mi/Mj)]0.5
, (C.50)
with
λtr,i
λtr,j
=
Γj
Γi
exp(0.0464Tr,i) − exp(−0.2412Tr,i)
exp(0.0464Tr,j) − exp(−0.2412Tr,j)
, (C.51)
where M is the molecular weight and Γ is deﬁned as
Γi = 210
Tc,iM3
i
P4
ci
(1/6)
, (C.52)
where T is in K, p is in bar, M is in g/mol and λ is in W/mK. This model yields an error
of less than 5% in the prediction of the thermal conductivity of the gas mixture.

C.3 Software packages 155
C.2.5 Surface tension of mixtures
Lucas and Luckas [92] in VDI-W¨armeatlas [157] recommended the following method for
calculation of the mixture surface tension
σm = p2/3
c,mT1/3
c,m
1 − tr,m
am
nm
bm , (C.53)
where
bi = 0.1196. 1 +
Ts,ri ln(pc,m/1.01325)
1 − Ts,ri
, bm = xibi , (C.54)
am = 1, nm = 11/9, Tc,m =
j=1
xiTc,j, υc,m =
j
xjυc,j, Zc,j =
pc,jυc,j
RTc,j
, (C.55)
Zm =
j
xjZc,j, pc,m =
RTc,mZc,m
υc,m
, Ts,ri =
Tb,i
Tc,i
, (C.56)
where Tb,i=T (p=1.01325 bar) is the normal boiling point temperature of the pure com-
ponent i. T is in K, p is in bar and σ is in N/m. The Lucas and Luckas correlation yields
an error of <5%.
C.3 Software packages
There exists a number of software packages for the prediction of thermodynamic and
transport properties. These include:
1. ASPEN Plus (
2. CHEMCAD
3. SUPERPRO
4. REFPROP

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