Heat Exchangers: Design, Operation ... - 123SeminarsOnly

Heat Exchangers: 

Design, Operation, Maintenance and Enhancement 

Ali A. Rabah (BSc., MSc., PhD., MSES) 

Department of chemical engineering, 

University of Khartoum, 

P.O. Box 321, 

Khartoum, Sudan 

Email: rabass@hotmail.com

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 



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 



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 



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 



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 demonstration 

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 procedure 

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 troubleshooting 

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 update. 

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 temperature, 

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 exchangers 

• 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, tubestubesheet 

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 selection, 

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, uncontrolled 

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 practices 

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 fundamentals 

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 networking 

break of 20 minutes during each morning and afternoon session. 

1.2 Instructor 

Faculty: Ali. Rabah, BSc. MSc., PhD., MSES., Assistant professor, Department 

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 sponsored 

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 software 

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 ≥ 700m 2 /m 3 ) 

2. non-compact (surface area density < 700m 2 /m 3 ) 

• 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 


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 services. 

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. 

Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : rabahss@hotamil.com 

13


– Chiller: one stream a process fluid being condensed at sub-atmospheric temperatures 

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 

Elbew 3/4" 

Valve 3/4" 

Galv. pipe 

Threaded 3/4" 

Bypass 

pump 

Part A 

Double Pipe Heat Exchanger 

Scale: None Sheet No.1 Date: 08.12.2003 

Designed by: Dr.-Ing. Ali A. Rabah 

Tee 2"x1/2" 

Union 2" 

Tee 3/4"x1/2" 

Part B 

Flanged Gland 2" 

Galv. pipe 2" 

Cu pipe 3/4" 

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 

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: 


Flow meter 

Bypass


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 m 2 (215 ft 2 ) 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 cleaning 

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. 

2.4.3 Floating head 

Figure 2.6. U-tube heat exchanger. 

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 advantage 

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 Utubes 

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 longitudinal 

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 alternating 

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 prevent 

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


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 heattransfer 

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 rectangular 

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 exchangers. 

The design thermal effectiveness is usually 60 % and below and the heat transfer area 

density is usually less than 300 m 2 m 3 . 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 nomenclature 

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 33 1/8 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 Tubular 

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. 



Packed Internal Outside 

Type Fixed lantern-ring floating head packed Pull-through 

of design tube sheet U-tube floating head (split backing ring) floating head floating head 

T.E.M.A. rear 

-head type L or M or N U W S P T 

Relative cost increases 

from A (least 

expensive) through 

E (most expensive) B A C E D E 

Provision for 

differential expansion Expansion Individual tubes Floating head Floating head Floating head Floating head 

joint in free to expand 

shell 

Removable bundle No Yes Yes Yes Yes Yes 

Replacement bundle 

possible No Yes Yes Yes Yes Yes 

Individual tubes 

replaceable Yes Only those in Yes Yes Yes Yes 

outside row 

Tube cleaning by 

chemicals inside 

and outside Yes Yes Yes Yes Yes Yes 

Interior tube 

cleaning mechanically Yes Special tools required Yes Yes Yes Yes 

Exterior tube 

cleaning mechanically: 

Triangular pitch No No No No No No 

Square pitch No Yes Yes Yes Yes Yes 

Hydraulic-jet 

cleaning: 

Tube interior Yes Special tools required Yes Yes Yes Yes 

Tube exterior No Yes Yes Yes Yes Yes 

Double tube 

sheet feasible Yes Yes No No Yes No 

Number of tube passes No practical Any even Limited to one No practical No practical No practical 

limitations number possible or two passes limitations limitations limitations 

Internal gaskets 

eliminated Yes Yes Yes No Yes No 


Table 3.2. Features of TEMA Shell-and-Tube-Type Exchangers.

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 pattern 

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 manufactured.) 

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 modified 

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 lightwall-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 floatingheat 

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 splitring 

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/in 2 ) for pipe shell exchangers and is 

limited to 1034 kPa (150 lbf/in 2 ) 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 Exchanger 

Manufacturers Association, 6th ed., 1978.) 

Outside-Packed Floating-Head Exchanger: (Fig. 3.8) The shell-side fluid is contained 

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 differential 

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 exchanger. 

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 slipon 

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 important 

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 increases 

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- (0.037in-) 

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 

Flow 

d o 

Square pitch 

p t 

3.4.5 Tube side passes 

p t 

Equilateral triangular pitch 

Figure 3.14. Tube patterns. 

d o 

Rotaed square 

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 

5 

2 

1 

2 

1 

6 

3 

4 

Two tube passes 

5 

2 

1 

2 

Four tube passes 

Six tube passes 

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 

4 

1 

6 

3 

4 

p t


3.5.2 Shell-Side Arrangements 

1. The one-pass shell (Fig. 3.1E) is the most commonly used arrangement. Condensers 

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 except 

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 

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 74d 0.75 (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 provided 

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 exchangers 

with a horizontal baffle cut and a horizontal pass rib for directing tube 


a 

b 

c 

d


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 impingement 

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 adequate 

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 tubeside 

nozzles when entrance velocity is high. 

(a) 

(c) 

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 bundles 

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. Arrangements 

to reduce tube-bundle bypassing include: 


(B) 

(d)


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. 

Rods 

3.6.6 Tubesheets 

Spacer 

baffle 

Figure 3.19. Baffle spacers and tie rods. 

Tube plate 

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. 



3 mm 

a 

0.4mm 

Figure 3.20. Tube sheet joint 

b c 

The tubesheet is in contact with both fluids and so must have corrosion resistance allowances 

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 additional 

tightness beyond that of tube rolling or may be strength-welded. Strengthwelded 

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 pressures. 

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 double- tube-sheet designs require special consideration (see Fig. 3.20c). 


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) 

where 

55 

Q = UA∆TlmF (4.1) 

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: 

T hi 

T ci 

T ci 

T hi 

T ho 

ΔT 1 ΔT2 

Cocurrent 

T co 

T co 

T ho 

Thi ΔT1 Tco T co 

T hi 

Counter current 

Tho ΔT2 T ci 

T ci 

T ho 

ΔT 1 

T hi 

T hi 

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 

T c 

Shell and Tube 

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 


T ho 

T ho 

T co 

T ci 

T ci 

T co

4.1 LMTD-Method 57 

Thi =110 C o 

Tci = 35 C o Tci = 35 C o 

T ci 

T hi 

Tho=75 C o 

ΔT 1 =75 

ΔT =10oC Cocurrent 

1. counter current (see Fig.4.2) 

2. co-current (see Fig.4.2) 

Tco =65 C o 

T co 

T ho 

2 

ΔT1 =45 C o 

Thi=110 C o 

Tco =65 C o 

T co 

T hi 

Counter current 

Figure 4.2. Temperature distribution 

∆Tlm = ∆T2 − ∆T1 

ln(∆T2/∆T1) 

∆Tlm = ∆T2 − ∆T1 

ln(∆T2/∆T1) 

4.1.2 Correction Factor 

• For double pipe heat exchanger 

Tho=75 C o 

ΔT C 

2=40 o 

Tci= 35 C o 

T ci 

T ho 

= 10 − 75 

ln(10/75) = 32.26o C (4.3) 

= 40 − 45 

ln(40/45) = 42.45o C (4.4) 

F = 1 (4.5) 

• Shell and tube heat exchanger. For a 1 shell 2 tube pass exchanger the correction 

factor is given by: 

 

(R 

F = 

2 + 1) ln 

1−S 

1−RS 

⎧ √ ⎫ 

⎨ 2−S R+1− (R2 +1) ⎬ 

(R − 1) ln √ 

⎩ 2−S R+1− (R2 +1) ⎭ 

(4.6) 

where 

or in words 

R = 

Range of shell fluid 

, S = 

Range of tube fluid 

R = T1 − T2 

, S = 

t2 − t1 

t2 − t1 

T1 − t1 

Range of tube fluid 

Maximum temperature difference 

(4.7) 

(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 

 

R2+1 ln 2(R−1) 

F = 

1−S 

1−RS 

ln 2/S−1−R+(2/S) 

√ √ (4.9) 

(1−S)(1−RS)+ R2 +1 

These equations are plotted on fig.4.4 

√ √ 

2/S−1−R+(2/S) (1−S)(1−RS)− R2 +1 



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 = 35 o C, T2 = 65 o C, t1 = 110 o C, t2 = 75 o C 

R = T1 − T2 

t2 − t1 

= 35 − 65 

From the graph of fig.4.4 

75 − 110 = 0.86, S = t2 − t1 

T1 − t1 

1. for 1-2 shell and tube heat exchanger F=0.92 

2. for 2-4 shell and tube heat exchanger F=0.98 

T 2 

t 1 

t 2 

= 75 − 110 

35 − 110 

1-2 Shell and Tube 

2-4 Shell and Tube 

T1 T1 t1 T 2 

= 0.467 (4.10) 

Figure 4.3. Temperature distribution for 1-2 and 2-4 shell and tube heat exchanger 


t 2

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 additional 

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 values 

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/m 2 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/m 2 o C 

ho = outside fluid film coefficient, W/m 2 o C 

hi = inside fluid film coefficient, W/m 2 o C 

hod = outside dirt coefficient (Fouling factor), W/m 2 o C 

hi = inside dirt coefficient,W/m 2 o C 

kw = thermal conductivity of the tube wall material, W/m o 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 

 

= f Re, P r, 

k d 

 

µ 

, (4.12) 

L µw 

and for natural convection as 

Nu = hd 

k 

= f 

 

Gr, P r, 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, organic 

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 

This is related to the heat exchanger size and capacity as 

Where NT U is number of transfer unit and is defined as 

(4.14) 

ε = f(NT U, C) (4.15) 

NT U = N = UA 

Cmin 

and C is the heat capacity ratio defined using energy equation as: 

(4.16) 

Q = MhCph(Thi − Tho) = McCpc(Tco − Tci) (4.17) 



Table 4.2. Fouling factor 

Fluid Coefficient (W/m 2 o C) Factor (resistance (m 2 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 

MhCpc > McCpc ⇒ Cmin = McCpc, Cmax = MhCph 

(4.18) 

(4.19) 

Qmax = Cmin(Thi − Tci) (4.20) 

C = Cmin 

Cmax 

εh = Thi − Tho 

, εc = 

Thi − Tci 

Tco − Tci 

Thi − Tci 

ε = ∆Tc 

Tspan 

where Tspan is defined in fig. 4.5 for counter current flow 

(4.21) 

(4.22) 

(4.23) 


4.2 ε- NTU 63 

T span 

Thi 

Tco 

Tho 

0 A 

Tci 

Figure 4.5. Temperature distribution in counter current flow 

The ε equation for various heat exchanger configuration is given as 

• Parallel flow 

• Counter current flow 

• Cross flow 

ε = 

ε = 

1. Both fluid unmixed mixed 

where 

2. Both fluid mixed 

ε = 

3. Cmax mixed, Cmin unmixed 

1 − exp [−N(1 + C)] 

1 + C 

1 − exp [−N(1 + C)] 

1 − C exp [−N(1 − C)] 

ε = 1 − exp 

 

exp(−NCn) − 1 

n = N −0.22 

Cn 

 

1 

1 − exp(−N) − 1 + 

−1 C 

1 

− 

1 − exp(−NC) − 1 N 

Δθ 

(4.24) 

(4.25) 

(4.26) 

(4.27) 

(4.28) 

ε = 1 

{1 − exp [−C (1 − exp(−N))]} (4.29) 

C 



4. Cmax unmixed, Cmin mixed 

• One shell pass, 2,4, 6 tube passes 

• Condenser 

• Evaporator 

 

ε = 1 − exp − 1 

 

[1 − exp(−NC)] 

C 

⎧ 

⎪⎨ 

ε = 2 1 + C + (1 + C 

⎪⎩ 2 

1 + exp −N (1 + C 

) 2 ) 

1 − exp 

−N (1 + C2 ) 

⎫ 

⎪⎬ 

⎪⎭ 

ε = 1 − e −N 

ε = 1 − e −N 

−1 

(4.30) 

(4.31) 

(4.32) 

(4.33) 

Alternatively these equations are presented in a graphical form. The various curves of ε 

vs NT U can be found in textbooks like Kern (1964( and Perry and Green (2000). 

4.3 Link between LMTD and NTU 

• Cocurrent 

• Counter current 

 

∆T1 

ln 

∆T2 

 

∆T1 

ln = ln 

∆T2 

 

Thi − Tci 

= ln 

= Nh + Nc 

Tho − Tco 

 

Thi − Tco 

Tho − Tci 

= Nh − Nc 

(4.34) 

(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 NT U methods by expressions 

Θ = ∆Tm 

Tspan 

= ε 

NT U 

(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


two value. Alternatively, the method suggested by Frank (1978) can be used; in which 

Q = A [U2(T1 − t2) − U2(T2 − t1)] 

ln (5.1) 

U2(T1−t2) 

U1(T2−t1) 

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/cm 2 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/cm 2 , with 0.1 

kg/cm 2 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. Viscosity 

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 considerations), 

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 availability 

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 fundamental 

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: 

or 

Nu = CRe a P r b 

h = C k 

D 

GD 

µ 

µ 

a Cpµ 

k 

µw 

c 

b µ 

, (5.2) 

µw 

c 

(5.3) 



where 

Nu = hde 

P r = 

k Nusselt number 

Cpµ 

Re 

de 

A 

k 

ρud 

µ 

4A 

P 

Prandtl number 

Reynolds number 

hydraulic diameter 

cross-sectional area 

P wetted perimeter 

u fluid velocity 

µw fluid viscosity at the tube wall temperature 

k fluid thermal conductivity 

fluid specific heat 

Cp 

⎧ 

⎪⎨ 

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 ∝ µ 0.8−0.33 = µ 0.47 

(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/hm 2◦ C, followed by hydrocarbon liquids (thermal conductivity 

between 0.08 and 0.12 kcal/hm ◦ C) at 250-1,300 kcal/hm 2◦ C, and then hydrocarbon gases 

(thermal conductivity between 0.02 and 0.03 kcal/hm ◦ C) at 50-500 kcal/hm 2◦ 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. 


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 4fG2 

= , 

d 2dρ 

(5.5) 

where G is the mass flux in kg/m2s and f is the friction factor calculated using a Blasiustype 

model as 

⎧ 

0.3164 

⎪⎨ Re 

f = 

⎪⎩ 

0.25 Re ≥ 2320 

64 Re < 2320 . 

Re 

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. Consequently, 

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 

m 2 . 

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/cm 2 per shell is permitted for liquid streams. If there are five 

such preheat exchangers, a total pressure drop of 3.5 kg/cm 2 for the circuit would be 

permitted. If the pressure drop through two of these exchangers turns out to be only 0.8 

kg/cm 2 , the balance of 2.7 kg/cm 2 would be available for the other three. 



5.4 Shell side design 

Shell side design The shellside calculations are far more complex than those for the tubeside. 

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 shellside 

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 employed 

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. 



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 circumstances, 

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 vibration. 

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 penetration 

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 heattransfer 

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 heattransfer 

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. Reducing 

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 designs. 

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. 



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 passpartition 

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 decreased, 

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 influencing 

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 



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 

where 

Nu = 0.36Re 0.55 P r 1/3 

µ 

µw 

0.14 

, (5.7) 

Nu = hde 

P r = 


Cpµ 

Re = 

k Prandtl number 

Gde 

de = 

µ Reynolds number 

4A 

A = 

P hydraulic diameter 

cross-sectional flow area 

P = wetted perimeter 

G = 

As = 

M 

As 

Mass flux 

(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 = 

⎪⎩ 


where 

L=tube length 

p 2 t −πd2 o /4 

πdo 

0.87p 2 t /2−πd2 o/8 

πdo/2 

∆p = 4f 

for square pitch 

for equilateral triangular pitch 

 

Ds 

d 

ρu2 L 

2 lb 

−0.14 µ 

µw 

⎧ 

⎪⎨ 

f = 

⎪⎩ 

0.3164 

Re0.25 Re ≥ 2320 

64 Re < 2320 . 

Re 

, (5.8) 

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 cc pc(Tc2-T c1)=MhC ph(Th1-Th2) Step2 

Calculate physical properties 

Step3 

Assume value of overall 

coefficient U o,ass 

Step 4 

Decide number of shell and 

tube passes 

Calculate ΔT , F and ΔT 

lm m 

Step 5 

Determine heat transfer area 

required Ao=q/Uo,assΔTm 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 

Set U o,ass =U o,cal 

No 

No 

Yes 

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 

U o,cal 

0

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 fabrication 

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 


7 Storage, Installation, Operation and Maintenance 

Proper storage, installation handling and correct start up emergency, and shutdown procedure 

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, certain 

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. 


83

84 7 Storage, Installation, Operation and Maintenance 

1. On receipt of the heat exchanger, inspect for shipping damage to all protective covers. 

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 applications 

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 atmosphere 

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 condensation 

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 surfaces 

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 information 

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 recoated 

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 lightbodied 

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 protective 

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. Oversized 

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 procedures 

for most applications. If in doubt, consult the nearest manufactuerer representative 

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 published 

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, depending 

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, exercise 

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. 


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 sleeving 

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 performing 

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, rebundling, 

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 


91

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 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 exchangers. 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.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 lowcost 

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 exchangers. 

In the tubesheet, tubes can be re-expanded to strengthen the original tubeto-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 removal, 

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 exchanger 

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% throughwall 

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 nondestructive 

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. 

Partial Length Structural Sleeve 

Figure 8.2. Full Length Sleeve Expansion 

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 withstand 

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% throughwall 

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 projectiles 

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 conditions 

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 calculations, 

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 sleeveto-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 operating 

environment. However, the production rates for sleeve installation are lower 

when full length expansions are performed. While full length expansion is typically 

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 

8.4.3 Tube Expansion 

Figure 8.4. Required Clearance for Sleeve Installation 

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 eliminated, 

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-totubesheet 

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 spraying 

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 exchangers 

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 requirements 

for the expansion. Before performing tube expansions into heat exchanger 

tubes, testing needs to be performed to set the tooling operating parameters. Depending 

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, qualification 

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 replacement. 

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 operations. 

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 characteristics 

• 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. Although 

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 economical 

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 overstressing 

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 expansion 

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 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. 

9.3.2 Cause of differential thermal expansion 

1. Unusual situation that lead to unexpected differential thermal expansion, for example, 

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 interactions 

between heating/cooling fluids and process fluids. For relatively low-pressure 

equipment ( 1000 psig), however, a complete failure should be considered 

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 thickness 

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 corrosion 

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 (Appendix 

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 suggests 

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 essentially 

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 remained 

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 150 o 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 developed 

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 resistance 

during the low temperature excursion caused by depressuring the acetylene converter 



system. The heat exchanger that failed was fabricated from ASTM A515 grade 70 carbon 

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 subsequent 

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 solutions 

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 conjunction 

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), 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 chemicals. 

For heat transfer problems involving highly reactive/ hazardous materials, a triplewall 

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. Alternatively 

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 separators 

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 between 

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 

Control 

vlave 

55 C o 

125 C o 

Water 

30 C o 

Organic 

Symptom: Shellside outlet 

temperaturee cannot be 

controlled within desired 

range (55-62 oC) 

by 

controlling flow of 125 C o 

water to tubes. The heat 

exchanger is 4 tube pass. 

55-62 C o 

uncontrolled 

9.9.2 Shell assumed banana-shape 

487 C o 

200 C o 

67 C o 

70 C o 

Water 

30 C o 

Organic 

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 

Control 

vlave 

Figure 9.5. Shell side temperature uncontrolled 

belows joint 

560 C o 

600 C o 

Figure 9.6. Shell assumed banana-shape 

55-62 C o 

controlled 

Bypass 

Cure: Tube side water 

temperature reduced to 70oC 

and control valve removed. 

Control valve is installed 

in new shellside bypass 

line 

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. 


9.9 Troubleshooting Examples 119 

9.9.3 Steam condenser performing below design capacity 

Steam 

Vent 

ondensate 

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 

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 extremely 

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 temperature. 

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 design 

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 



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A Heat transfer coefficient 

A.1 Single phase 

A.1.1 Inside tube: Turbulent flow 

where 

Nu = CRe a P r b 

µ 

µw 

c 

131 

, (A.1) 

Nu = hde 

P r = 


Cpµ 

Re 

de 

A 

k 

ρud 

µ 

4A 

P 

Prandtl number 

Reynolds number 

hydraulic diameter 

cross-sectional area 

P wetted perimeter 

u fluid velocity 

µw fluid viscosity at the tube wall temperature 

k fluid thermal conductivity 

fluid specific heat 

Cp 

⎧ 

⎪⎨ 

C = 

⎪⎩ 

A.1.2 Inside tube: Laminar flow 

A.1.3 Shell side 

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 

 

Nu = 1.86 ReP r d 

1/3 0.14 µ 

, (A.2) 

L µw 

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 

where 

Nu = 0.36Re 0.55 P r 1/3 

µ 

µw 

0.14 

, (A.3) 

Nu = hde 

P r = 


Cpµ 

Re = 

k Prandtl number 

Gde 

de = 

µ Reynolds number 

4A 

A = 

P hydraulic diameter 

cross-sectional flow area 


G = 

As = 

M 

As 

Mass flux 

(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) 

d o 

Square pitch 

⎧ 

⎪⎨ 

de = 

⎪⎩ 

p 2 t −πd2 o/4 

πdo 

0.87p 2 t /2−πd2 o/8 

πdo/2 

p t 

for square pitch 

for equilateral triangular pitch 

p t 

Equilateral triangular pitch 

Figure A.1. Tube arrangement 

A s 

Cross-flow area 


A.2 Condensation 133 

A.1.4 Plate heat exchanger 

where 

Nu = 0.26Re 0.65 P r 0.4 

µ 

µw 

0.14 

, (A.4) 

Nu = hde 

P r = 

= k Nusselt number 

Cpµ 

Re = 

= k Prandtl number 

ρupde Gde = µ µ = Reynolds number 

de = hydraulic diameter, taken as twice the gap between the plates 

A = cross-sectional flow area 


G = 

Af = 

M 

Af 

= Mass flux 

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 

where 

hm = 0.943 

 

3 1/4 

k ρ∆ρgλ 

µ∆T L 

, (A.5) 

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 

where 

hm = 0.725 

k 3 ρ∆ρgλ 

µ∆T do 

1/4 

do = out side diamter of the tube 

A.2.3 Condensation on banks of horizontal tube 

hm = 0.725 


µ∆T Jdo 

1/4 

, (A.6) 

, (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 

A.3 Two phase flow: Pure fluid 

A.3.1 Steiner [140] correlation 


µ∆T d 

1/4 

, (A.10) 

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 = 

h 3 n + h 3 1/3 c 

, (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 ˙x 

⎩ 

0.4 (1 − ˙x) 0.01 

⎣ hG0 

hL0 

⎛ 

˙x 0.01 ⎝1 + 8(1 − ˙x) 0.7 

⎤ 

0.37 

ρL 

ρG 

⎦ + 

⎞⎤ 

0.67 −2 

ρL ⎠⎦ 

ρG 

⎫−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)P r 

1 + 12.7(ξ/8) 0.5 (P r 2/3 − 1) 

, (A.13) 

taken in to account the respective dimensionless group NuL0, NuG0, ReL0, ReG0, P rl and 

P rg. These dimensionless groups are defined as 

NuL0/G0 = hL0/G0d 

kL/G 

, (A.14) 


A.3 Two phase flow: Pure fluid 135 

The friction factor is 

ReL0/G0 = ˙md 

µL/G 

, (A.15) 

P rL/G = µL/Gcp,L/G 

. (A.16) 

kL/G 

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) 

 

where ϕ is the stratified angle. The Reynolds number is given as 

and 

ReG = 

˙m ˙xdhyd 

ɛµG 

hG = NuGkG 

dhyd 

, (A.19) 

, (A.20) 

. (A.21) 

The void fraction is calculated using the Rauhani [117] model given as 

ε = ˙x 

ρG 

 

 

˙x 

(1 + 0.12(1 − ˙x)) 

ρG 

+ 1 − ˙x 

 

ρL 

+ 1.18(1 − ˙x)[gσ(ρL − ρG)] 1/4 

˙mρ . 1/2 

−1 L 

The wetting boundary can be estimated (see Fig. A.2) from the void fraction as 

ε = fG 

fG + fL 

(A.22) 

. (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 

where ϕG = 0.5ϕ. 

Φ = ϕG 

2π 

, (A.25) 



f L 

f G 

U L 

U G 

U i 

ϕ 

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 

n(pr) hnb ˙q 

= ψCf F (pr)F (Ra)F (d)F ( ˙m, ˙x) . (A.26) 

ho ˙qo 

The value with a subscript ”o” is a reference value. 

The pressure function is given as 

F (pr) = 2.692p 0.43 

 

6.5 1.6pr r + 

1 − p4.4 

r 

and the mass flux function is given as 

where 

F ( ˙m, ˙x) = ˙m 

˙mo 

0.25 ⎡ 

⎣1 − p 0.1 

r 

˙q 

h 

qcr,nb 

d 

0.3 

, (A.27) 

⎤ 

˙x ⎦ , (A.28) 

˙qcr,cb = 2.79 ˙qcr,0,1p 0.4 

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,0ρ 0.5 

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 − 0.3p 0.3 

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/m 2 K W/m 2 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 = 0.023Re 0.8 

G P 0.4 kG 

rG 

d 

, (A.34) 

with Reynold number given as 

˙m ˙xd 

ReG = , (A.35) 

εµG 

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 = 55p 0.12 

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 

The liquid Reynolds number is given as 

hc = 0.0133Re 0.69 

L P 0.4 kL 

rL 

d 

ReL = 

4 ˙m(1 − ˙x)δ 

(1 − ε)µG 

where δ is the liquid film thickness it is given as 

δ = 

πd(1 − ε) 

2(2π − ϕdry) 

. (A.38) 

. (A.39) 

, (A.40) 



where ϕdry is 

ϕdry = ϕstrat 

( ˙mwavy − ˙m) 

, (A.41) 

( ˙mwavy − ˙mstrat) 

where ϕstrat is calculated iteratively from equation A.24. The mass flux under a stratified 

and wavy flow pattern is 

and 

˙mwavy = 

 

˙mstrat = (226.3)2 fLf 2 GρG(ρL − ρG)µLg cos Θ 

0.3164(1 − ˙x) 1.75 π 2 µ 0.25 

L 

16f 3 GgdρLρG 

˙x 2π2 (1 − (2hL − 1) 2 ) 0.5 

 

2 π 

25h2 (1 − x) 

L 

F1( ˙q) 

× 

, (A.42) 

F2( ˙q) 

W e 

+ 

F r L 

1 

0.5 

+ 50 , 

cos Θ 

(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 

2 

+ 64.8 

˙q 

 

; 

 

F2( ˙q) = 18.8 

˙q 

 

+ 1.023 . (A.44) 

˙qcrit 

˙qcrit 

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 

˙qcrit 

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 

and 

hn = 0.6683Co −0.2 (1 − ˙x) 0.8 hL0f(FrL0) + 1058.0Bo 0.7 (1 − ˙x) 0.8 FF lhL0 , (A.46) 

hc = 1.136Co −0.9 (1 − ˙x) 0.8 hL0f(FrL0) + 667.2Bo 0.7 (1 − ˙x) 0.8 FF lhL0 , (A.47) 

respectively, where F rL0 is the liquid Froude number, Bo is the boiling number and Co 

is the convection number. These dimensionless groups are defined as 

F rL0 = ˙m 

, Bo = 

ρLgd 

The function f(FrL0) is defined as 

˙q 

, Co = 

˙m∆hv 

ρG 

ρL 

0.5 1 − ˙x 

f(FrL0) = (25FrL0) 0.324 FrL0 < 0.04 , 

f(FrL0) = 1 FrL0 ≥ 0.04 , 

˙x 

0.8 

. (A.48) 

where FF l 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 10 4 ≤ ReL0 ≤ 5 × 10 6 and it is given as 

NuL0 = hL0d 

k = 

ReL0P rL(ξ/2) 

1.07 + 12.7(P 2/3 

. (A.49) 

rL − 1)(ξ/2) 0.5 

The Gnielinski [46] correlation (equation A.13) is valid in the range of 0.5 ≤ PrL ≤ 

2000 and 2300 ≤ ReL0 ≤ 5 × 10 4 . 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 microconvective 

(or nucleate boiling) portion hn and b) a macro-convective (or forced convective) 

portion hc as 

h = hcF + hnS , (A.50) 

where hc is calculated using the Dittus and Boelter [33] correlation as 

where 

hc = 0.023 kL 

d Re0.8 L P r 0.4 

L , (A.51) 

(1 − ˙x) ˙md 

ReL = , P rL = cpLµL 

, (A.52) 

µL 

The suppression factor for the convection part is 

⎧ 

⎪⎨ 

F = 

⎪⎩ 

kL 

1 if 1/Xtt > 0.1 

2.35 1 

Xtt + 0.213 0.736 

and the Martinelli parameter Xtt is given as 

 

1 − ˙x 

X = 

˙x 

The nucleate boiling heat transfer coefficient is 

where 

hn = 0.00122 

if 1/Xtt ≤ 0.1 

0.875 0.5 0.125 ρG µL 

ρL 

k0.79 L c0.45 p,L ρ0.49 L 

σ0.5 µ 0.29 

L ρ0.24 V 

G ∆h 0.24 

µG 

, 

. (A.53) 

∆T 0.24 

sat ∆p 0.75 

sat , (A.54) 

∆Tsat = Tw − Ts; ∆psat = p(Tw) − p(Ts); Retp = ReLF 1.25 . (A.55) 

The suppression factor for the nucleate part is 

S = 

1 

1 + 2.53 × 10 −6 Retp 

. (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] correlation 

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 (0.1−2F rL) 

)F r L if F r < 0.05 

1 + 24, 000Bo 1.16 + 1.37(1/xtt) 0.86 if F r ≥ 0.05 

and the suppression factor for the nucleate part is 

⎧ 

⎪⎨ 

S = 

⎪⎩ 

(1 + 0.00000115F 2 ReL) −1 F r 1/2 

L if F r < 0.05 

(1 + 0.00000115F 2 ReL) −1 if F r ≥ 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 

where 

⎧ 

⎪⎨ 

N = 

⎪⎩ 

hc = 1.8hLN −0.8 , (A.58) 

Co F rL > 0.04 

0.38F r −0.4 

L Co F rL < 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 

where 

• For N > 1 

• For 1 > N > 0.1 

• For N < 0.1 

⎧ 

⎪⎨ 230hLBo 

hn = 

⎪⎩ 

0.5 Bo > 0.0003 

1 + 46hLBo0.5 Bo < 0.0003 

hn = F hLBo 0.5 exp(2.74N −0.1 ) . (A.59) 

hn = F hLBo 0.5 exp(2.47N −0.15 ) , (A.60) 

⎧ 

⎪⎨ 14.7 Bo > 0.0011 

F = 

⎪⎩ 

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 10 4 

1 

× Bo + 1.5 

Xtt 

2/3 0.6 

where hL is calculated using Dittus-Boelter [33] correlation equation A.51. 

A.3.8 Dembi et al. [30] correlation 

, (A.61) 

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 

and 

hn = 23388.5 kL 

d 

 

hc = 0.115 kL 

d 

˙q 

ρG∆hV ϖ 

respectively. The parameter ϖ is defined as 

A.3.9 Klimenko [84] correlation 

0.64 gd 

∆hV 

0.27 

˙m 2 d 

ρL∆hV ϖ 

0.14 

, (A.62) 

 

˙x 4 (1 − ˙x) 2 

0.11 

2 0.14 

˙m ∆hV 

P 

ρLgσ 

0.27 

rL , (A.63) 

ϖ = 0.36 × 10 −3 p −1.4 

r . (A.64) 

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 

where 

hn = 

 

hn1 = 7.4 × 10 −3 

hn2 = 0.087 kL 

b 

hn1 NCB < 1.6 × 10 4 

hn2 NCB > 1.6 × 10 4 , 

kw 

kw 

kL 

kL 

0.15 

P e 0.6 K 0.5 

p P r −1/3 

L , (A.65) 

0.09 

Re 0.6 

m 

ρG 

ρL 

0.2 

P r 1/6 

L , (A.66) 

 

 

 

qb 

p 

2σ 

P e = 

, Kp = 

, b = 

, (A.67) 

∆hV ρGaL 

σg(ρL − ρG) g(ρL − ρG) 

Rem = wmb 

, wm = ˙m 

 

1 + x 

νL 

ρL 

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 

where 

hn = 207 kL 

0.745 0.581 ˙q(b.d) ρG 

P 

b.d 

0.533 

rL , (A.69) 

⎧ 

⎪⎨ 

S = 

⎪⎩ 

kLTs 

(b.d) = 0.511 

A.4 Two phase flow: Mixture 

A.4.1 Steiner [140] correlation 

 

2σ 

ρL 

g(ρL − ρG) 

 

F = 2.37 0.29 + 1 

 

Xtt 

0.5 

4048X 1.22 

tt Bo 1.13 Xtt < 1 

2.0 − 0.1X −0.28 

tt Bo −0.33 1 ≤ Xtt ≤ 5 

, (A.70) 

, (A.71) 

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 × 10 5 

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 = 5.10 3 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 

where 

h = 1.136Co −0.9 (1 − ˙x) 0.8 h L0 f(FrL0) + 667.2Bo 0.7 (1 − ˙x) 0.8 FF lhL0FD , (A.77) 

V1 = 

cpL 

∆hV 

a 

D12 

FD = 0.678 

1 + V1 

A.4.3 Bennett and Chen [8] correlation 

 

0.5 dT 

|y − x| 

dx 

, (A.78) 

. (A.79) 

Bennett and Chen [8] has extended the Chen [19] correlation (equation A.50) for mixture. 

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 nucleate 

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 

where 

and 

Fn = 

1 

2 , (A.83) 

{[1 + (b2 + b3)(1 + b4)](1 + b5)} 

 

1.01 − x1 

b2 = (1 − x1) ln 

+ x1 ln 

1.01 − y1 

⎧ 

⎪⎨ 

b3 = 

⎪⎩ 

0.1 x1 

y1 

 

x1 

y1 

0 x1 ≥ 0.01 

 

p 

b4 = 152 

− 1 x1 < 0.01 

pc,1 

0.66 

b5 = 0.92|y1 − x1| 0.001 

 

p 

pc,1 

+ |y1 − x1| 1.5 , (A.84) 

, 

, (A.85) 

0.66 

x1 

y1 = 1 for x1 = y1 = 0 , 

, (A.86) 

x1 and y1 is the liquid and vapor mole fraction of the more volatile component respectively. 

p and pc,1 is system pressure and critical pressure of the more volatile component 

respectively. 


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 

145 

, (B.1) 

where ˙m is the mass flux in kg/m2s and f is the friction factor calculated using a Blasiustype 

model as 

⎧ 

⎪⎨ 

f = 

⎪⎩ 

0.3164 

Re0.25 64 

Re 

Re 

Re 

≥ 2320 

< 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 heterogenous 

model. In the homogenous model the static pressure drop due to acceleration 

is 

 

 

dp 

2 d 1 

− = ˙m ˙x − 

dz dz ρL 

a 

1 

 

+ 

ρG 

1 

 

. (B.4) 

ρL 

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 

 

4 ˙q ˙m 

∆pa = 1 − 

d∆hvρG 

ρG 

 

∆L . 

ρL 

(B.6) 

In the separated flow model the static pressure drop due to acceleration can be derived 

from the momentum balance as 

 

dp 

− 

dz 

 

2 

2 d ˙x 

= ˙m 

dz ερG 

+ (1 − ˙x)2 

 

(1 − ε)ρL 

. (B.7) 

a 

Integration of equation B.7 between the inlet i and outlet o of the test section yields 

−∆pa = −(po − pi)a = ˙m 2 

 

˙x 2 2 

εoρG,o 

+ (1 − ˙xo) 2 

(1 − εo)ρL,o 

− ˙x2 

2 

i (1 − ˙xi) 

− . 

εiρG,i (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 

 

 

˙x 

(1 + 0.12(1 − ˙x)) 

ρG 

+ 1 − ˙x 

 

ρL 

+ 1.18(1 − ˙x)[gσ(ρL − ρG)] 1/4 

˙mρ 1/2 

−1 L 

, (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 = 

⎪⎩ 

and the homogenous densityρH is given as 

1 

ρH 

The two phase Reynolds number Re is 

0.3164 

Re0.25 Re ≥ 2320 

64 Re < 2320 . 

Re 

= 1 − ˙x 

ρL 

Re = ˙md 

ηT P 

+ ˙x 

ρG 

. (B.11) 

, (B.12) 

where ηT P 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 

ηT P 

= 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 

ΨG/L , (B.14) 

dz f,L/G 

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 

where 

H = 

ΨL0 = E + 

3.24F H 

F r0.045 , (B.15) 

W e0.035 E = (1 − ˙x) 2 2 ρLfG0 

+ ˙x 

ρGfL0 

, (B.16) 

F = ˙x 0.78 (1 − ˙x) 0.24 , (B.17) 

0.91 0.19 

ρL µG 

ρG 

µL 

F r = ˙m2 

gdρ 2 H 

1 − µG 

µL 

0.7 

, (B.18) 

, (B.19) 

W e = ˙m2 d 

, (B.20) 

σρH 

d is tube diameter, σ is the surface tension and ϱH 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 

Re 0.25 

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 

ψ 2 L = 1 + C 

X 

1 

+ , (B.22) 

X2 ψ 2 G = 1 + C.X + X 2 , (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

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 

where 

ΨL0 = 1 + (Y 2 − 1) 

B ˙x (2/n−1) (1 − ˙x) (2/n−1) + ˙x 1−n 

Y 2 = (dpf/dz)G0 

(dpf/dz)L0 

n is 0.25 for a Blasius model. The parameter B is given by 

, (B.24) 

, (B.25) 

B = 55 

˙m 1/2 0 < Y < 9.5 , (B.26) 

B = 520 

Y ˙m 1/2 9.5 < Y < 28 , (B.27) 

B = 15000 

Y 2 ˙m 1/2 28 < Y . (B.28) 

The range of the validity of the Chisholm [22] correlation is µL/µG > 1000 and ˙m > 100 

kg/m 2 s. 


C Physical properties 

The fluid physical properties required for heat exchanger design are divided in thermodynamic 

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 demsity, 

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 = CP V AP A + (CP V AP B)T + (CP V AP C)T 2 + (CP V AP D)T 3 

149 

(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 = ANT A − 

ANT B 

T + ANT C 

(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: 

 

1 

log µ = V ISA 

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 procedure 

for the calculation of the vapor viscosity. 

for Tr ≤ 1 and pr ≤ ps/pc 

with 

η = (ηξ) r 1 

FpFQ 

ξ 

, (C.4) 

(ηξ) r = 0.600 + 0.760p α r + (6.990p β r − 0.6)(1 − Tr) , (C.5) 

α = 3.262 + 14.98p 5.508 

r and β = 1.390 + 5.746pr , (C.6) 


150 C Physical properties 

for 1≤ Tr ≤ 40 and 0≤ pr ≤ 100 

(ηξ) r = (η o ξ) 

 

1 + 

where η o is the low pressure viscosity given as 

Ap E r 

Bp F r + (1 + Cp D r ) −1 

 

, (C.7) 

η o ξ = [0.807T 0.618 

r − 0.357 exp(−0.449Tr) + 0.340 exp(−4.058Tr) + 0.018]F o p F o 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 

exp(a2T 

Tr 

γ r ) , (C.10) 

B = A(b1Tr − b2) , (C.11) 

C = c1 

D = d1 

exp(c2T 

Tr 

δ r ) , (C.12) 

exp(d2T 

Tr 

ɛ r ) , (C.13) 

E = 1.3088 , (C.14) 

F = f1 exp(f2T ς 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 ɛ -7.6351 

Fp = 1 + (F o p − 1) 

(ηξ) r 

η o ξ 

−3 

, (C.16) 

and 

FQ = 1 + (F o 

r (ηξ) 

Q − 1) 

ηo −1 

r (ηξ) 

− 0.007 ln 

ξ 

ηo 4 ξ 

, (C.17) 

where F o p and F o Q is low-pressure polarity and quantum factors respectively. These factors 

are 

F o p = 1 , 0 ≤ µr < 0.022 , (C.18) 

F o p = 1 + 30.55(0.292 − Zc) 1.7 , 0.022 ≤ µr < 0.075 , (C.19) 

F o 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 F o Q = 1.0 for all substances other than 

He, H2 and D2 . The reduced dipole moment µr is given as 

µr = µ2pc , (C.21) 

(kTc) 2 

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 

 

0.021357 

ηo(T ) = 100 

0.29572 

( MT 1/2 ) 

, (C.23) 

exp(Ω) 

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 

ɛ 

 

4 

 

+ C(n) log 

n=3 

kT 

ɛ 

n 

where ɛ/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.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 

where 

η1(T ) = Fv(T )ηo(T )ρ , (C.25) 

⎧ 

⎪⎨ 

 

13 

Fv(T ) = C A(1) + A(i) log 

⎪⎩ 

i=2 

kT 

ɛ 

⎫ 

−(i−1) ⎪⎬ 2 

⎪⎭ 

, (C.26) 

where C=0.6022137/0.2957 3 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 -0.17999496×10 1 2 0.466692621×10 2 

3 -0.53460794×10 3 4 0.33604074×10 4 

5 -0.13019164×10 5 6 0.33414230×10 5 

7 -0.58711743×10 5 8 0.71426686×10 5 

9 -0.59834012×10 5 10 0.33652741×10 5 

11 -0.12027350×10 5 12 0.24348205×10 4 

13 -0.120807957×10 3 

The third term in the viscosity equation C.22 is the contribution of the dense gas 

3 

η2(ρ, T ) = F (i, T )ρ 

i=1 

i+1 , (C.27) 



where 

⎧ 

⎪⎨ 

F (i, T ) = 

⎪⎩ 

1 0.219664285 2 ɛ − 0.83651107 × 10 kT 

−1 4 ɛ 

kT 

2 0.17366936 × 10−2 − 0.83651107 × 10−2 

ɛ 

kT 

3 0.167668649 × 10−3 2 ɛ − 0.149710093 × 10 kT 

−3 3 ɛ + kT 

0.77012274 × 10−4 4 ɛ 

kT 

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 correlation 

for the calculation of the surface tension 

m 

1 − Tr 

b , (C.28) 

σ = p 2/3 

c T 1/3 

c 

where the reduced pressure and temperature are defined as 

respectively. 

For a polar fluid like R134a the following quantities are valid 

a 

pr = p 

, Tr = T 

, , (C.29) 

pc 

a = 1 , (C.30) 

b = 0.1574 + 0.359ω − 1.769X − 13.69X 2 − 0.510ω 2 + 1.298ωX , (C.31) 

m = 1.210 + 0.5385ω − 14.61X − 32.07X 2 − 1.656ω 2 + 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 

ρ 1/3 

Cp . (C.34) 

M 

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 

Tc 

 

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/m 2 ) 


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 

where z is the compressibilty factor defined as 

and 

z 3 + z 2 + z(B 2 + B − A) = 0 . (C.36) 

h = ho + RT + 

z = pv 

RT 

A = aP 

R2 bp 

, B = 

T 2.5 RT 

 

v 

C.2 Physical properties: Mixture 

C.2.1 Liquid dynamic viscosity of mixtures 

0 

T 

. (C.37) 

. (C.38) 

dP 

R2T 2.5 

− p dv . (C.39) 

dT 

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 

n 

ln ηm = xi. ln ηL,i + 2.x1.x2.G12 , (C.40) 

i=1 

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 = η o m + ∆η , (C.41) 

where η o m is the mixture gas viscosity at a low pressure and ∆η is a correction factor for 

the high pressure viscosity 

η o n yiηG,i 

m = nj=1 yiφij 

i=1 

, (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 

φji = ηG,j 

ηG,i 

Mj 

, (C.43) 

φij . (C.44) 

Mi 



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 

The pseudo critical properties of the mixture are calculated as 

Tc,m = 

yjTc,j, υc,m = 

j=1 

j=1 

j 

yjυc,j, 

υc,m 

Zc,j 

= pc,jυc,j 

, 

RTc,j 

υc,m 

. (C.45) 

Zm 

= 

yj Zc,j, (C.46) 

Mm = 

yj Mm/1000 

Mj, ρc,m = , ρr,m = ρm 

, pc,m = RTc,m Zc,m 

, (C.47) 

where T is in K, p is in Mpa, υc,m is in m 3 /kmol, ρr,m is in kg/m 3 , 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 = 

j 

υc,m 

2 

XiλL,i − 0.72X1 X2|λL,2 − λL,1| , (C.48) 

i=1 

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 

nj=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 

with 

λtr,i 

λtr,j 

Aij = 

 

1 + (λtr,i/λtr,j) 0.5 ( Mj/ Mi) 0.25 2 

[8(1 + Mi/ Mj)] 0.5 

= Γj exp(0.0464Tr,i) − exp(−0.2412Tr,i) 

Γi exp(0.0464Tr,j) − exp(−0.2412Tr,j) 

where M is the molecular weight and Γ is defined as 

Γi = 210 

Tc,i M 3 i 

P 4 ci 

(1/6) 

, (C.50) 

, (C.51) 

, (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-Wärmeatlas [157] recommended the following method for 

calculation of the mixture surface tension 

nm 1 − tr,m 

bm , (C.53) 

where 

bi = 0.1196. 

σm = p 2/3 

c,mT 1/3 

c,m 

 

am 

1 + Ts,ri ln(pc,m/1.01325) 

1 − Ts,ri 

j=1 

j 

 

, bm = xibi , (C.54) 

am = 1, nm = 11/9, Tc,m = 

xiTc,j, υc,m = 

xjυc,j, Zc,j 

= pc,jυc,j 

, (C.55) 

j 

υc,m 

RTc,j 

Zm = 

xj Zc,j, pc,m = RTc,m Zc,m 

, Ts,ri = Tb,i 

, (C.56) 

where Tb,i=T (p=1.01325 bar) is the normal boiling point temperature of the pure component 

i. T is in K, p is in bar and σ is in N/m. The Lucas and Luckas correlation yields 

an error of

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