heat exchanger

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: [email protected]

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 : [email protected]

Table of contents 3

3.6.2 Baffle . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 48

3.6.3 Vapor Distribution . . . . . . . . . . . . . . . . . . . . . . . . . . . 51

3.6.4 Tube-Bundle Bypassing . . . . . . . . . . . . . . . . . . . . . . . . 51

3.6.5 Tie Rods and Spacers . . . . . . . . . . . . . . . . . . . . . . . . . . 52

3.6.6 Tubesheets . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 52

4 Basic Design Equations of Heat Exchangers 55

4.1 LMTD-Method . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 55

4.1.1 Logarithmic mean temperature different . . . . . . . . . . . . . . . 56

4.1.2 Correction Factor . . . . . . . . . . . . . . . . . . . . . . . . . . . . 57

4.1.3 Overall heat transfer coefficient . . . . . . . . . . . . . . . . . . . . 59

4.1.4 Heat transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . . 61

4.1.5 Fouling factor (hid, hod) . . . . . . . . . . . . . . . . . . . . . . . . . 61

4.2 ε- NTU . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 61

4.3 Link between LMTD and NTU . . . . . . . . . . . . . . . . . . . . . . . . 64

4.4 The Theta Method . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 64

5 Thermal Design 66

5.1 Design Consideration . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 66

5.1.1 Fluid Stream Allocations . . . . . . . . . . . . . . . . . . . . . . . . 66

5.1.2 Shell and tube velocity . . . . . . . . . . . . . . . . . . . . . . . . . 66

5.1.3 Stream temperature . . . . . . . . . . . . . . . . . . . . . . . . . . 67

5.1.4 Pressure drop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 67

5.1.5 Fluid physical properties . . . . . . . . . . . . . . . . . . . . . . . . 67

5.2 Design data . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 68

5.3 Tubeside design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 69

5.3.1 Heat-transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . . 69

5.3.2 Pressure drop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 70

5.4 Shell side design . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

5.4.1 Shell configuration . . . . . . . . . . . . . . . . . . . . . . . . . . . 72

5.4.2 Tube layout patterns . . . . . . . . . . . . . . . . . . . . . . . . . . 73

5.4.3 Tube pitch . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 73

5.4.4 Baffling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 74

5.4.5 Equalize cross-flow and window velocities . . . . . . . . . . . . . . . 76

5.4.6 Shellside stream analysis (Flow pattern) . . . . . . . . . . . . . . . 76

5.4.7 Heat transfer coefficient and pressure drop . . . . . . . . . . . . . . 77

5.4.8 Heat transfer coefficient . . . . . . . . . . . . . . . . . . . . . . . . 78

5.4.9 Pressure drop . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 78

5.5 Design Algorithm . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 79

6 Specification sheet 80


4 Table of contents

6.1 Information included . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80

6.2 Information not included . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80

6.3 Operation conditions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 80

6.4 Bid evaluation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 81

6.4.1 Factor to be consider . . . . . . . . . . . . . . . . . . . . . . . . . . 81

7 Storage, Installation, Operation and Maintenance 83

7.1 Storage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 83

7.2 Installation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 85

7.2.1 Installation Planning . . . . . . . . . . . . . . . . . . . . . . . . . . 85

7.2.2 Installation at Jobsite . . . . . . . . . . . . . . . . . . . . . . . . . 86

7.3 Operation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 87

8 Heat exchanger tube side mainenance (Repair vs replacement 91

8.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 91

8.2 Repair vs. Replace - Factors To Consider . . . . . . . . . . . . . . . . . . . 92

8.3 Heat Exchanger maintenance Options . . . . . . . . . . . . . . . . . . . . . 93

8.4 Repair option . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94

8.4.1 Plug . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 94

8.4.2 Sleeving . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 95

8.4.3 Tube Expansion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 100

8.5 Replacement option . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 103

8.5.1 Retubing . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 103

8.5.2 Rebundling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 104

8.5.3 Complete replacement (New unit) . . . . . . . . . . . . . . . . . . . 104

8.6 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 105

9 Troubleshooting 106

9.1 Heat exchangers’ problems . . . . . . . . . . . . . . . . . . . . . . . . . . . 106

9.2 Fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 106

9.2.1 Costs of fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 106

9.2.2 Facts about fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . 107

9.2.3 Types of Fouling . . . . . . . . . . . . . . . . . . . . . . . . . . . . 107

9.2.4 Fouling Mechanisms . . . . . . . . . . . . . . . . . . . . . . . . . . 107

9.2.5 Conditions Influencing Fouling . . . . . . . . . . . . . . . . . . . . . 107

9.2.6 Fouling control . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 108

9.2.7 Fouling cleaning methods . . . . . . . . . . . . . . . . . . . . . . . 108

9.3 Leakage/Rupture of the Heat Transfer Surface . . . . . . . . . . . . . . . . 109

9.3.1 Cost of leakage . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 109

9.3.2 Cause of differential thermal expansion . . . . . . . . . . . . . . . . 109

9.4 Corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110


Table of contents 5

9.4.1 Corrosion effects . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110

9.4.2 Causes of corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . 110

9.4.3 Type of corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110

9.4.4 Stress corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 110

9.4.5 Galvanic corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . 110

9.4.6 Pitting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111

9.4.7 Uniform or rust corrosion . . . . . . . . . . . . . . . . . . . . . . . 111

9.4.8 Crevice corrosion . . . . . . . . . . . . . . . . . . . . . . . . . . . . 111

9.4.9 Materials of Construction . . . . . . . . . . . . . . . . . . . . . . . 112

9.4.10 Fabrication . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 112

9.5 Troubleshooting . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 112

9.6 Past failure incidents . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 113

9.6.1 Ethylene Oxide Redistillation Column Explosion: . . . . . . . . . . 113

9.6.2 Brittle Fracture of a Heat Exchanger . . . . . . . . . . . . . . . . . 113

9.6.3 Cold Box Explosion . . . . . . . . . . . . . . . . . . . . . . . . . . . 114

9.7 Failure scenarios and design solutions . . . . . . . . . . . . . . . . . . . . . 114

9.8 Discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 116

9.8.1 Use of Potential Design Solutions Table . . . . . . . . . . . . . . . . 116

9.8.2 Special Considerations . . . . . . . . . . . . . . . . . . . . . . . . . 117

9.9 Troubleshooting Examples . . . . . . . . . . . . . . . . . . . . . . . . . . . 118

9.9.1 Shell side temperature uncontrolled . . . . . . . . . . . . . . . . . . 118

9.9.2 Shell assumed banana-shape . . . . . . . . . . . . . . . . . . . . . . 118

9.9.3 Steam condenser performing below design capacity . . . . . . . . . 119

9.9.4 Steam heat exchanger flooded . . . . . . . . . . . . . . . . . . . . . 119

10 Unresolved problems in the heat exchangers design 120

10.1 Future trend . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 120

Bibliography 121

A Heat transfer coefficient 131

A.1 Single phase . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 131

A.1.1 Inside tube: Turbulent flow . . . . . . . . . . . . . . . . . . . . . . 131

A.1.2 Inside tube: Laminar flow . . . . . . . . . . . . . . . . . . . . . . . 131

A.1.3 Shell side . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 131

A.1.4 Plate heat exchanger . . . . . . . . . . . . . . . . . . . . . . . . . . 133

A.2 Condensation . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 133

A.2.1 Condensation on vertical plate or outside vertical tube . . . . . . . 133

A.2.2 Condensation on external horizontal tube . . . . . . . . . . . . . . 133

A.2.3 Condensation on banks of horizontal tube . . . . . . . . . . . . . . 133

A.2.4 Condensation inside horizontal tube . . . . . . . . . . . . . . . . . . 134


6 Table of contents

A.3 Two phase flow: Pure fluid . . . . . . . . . . . . . . . . . . . . . . . . . . . 134

A.3.1 Steiner [140] correlation . . . . . . . . . . . . . . . . . . . . . . . . 134

A.3.2 Kattan et al. [77] correlation . . . . . . . . . . . . . . . . . . . . . . 137

A.3.3 Kandlikar [70] correlation . . . . . . . . . . . . . . . . . . . . . . . 138

A.3.4 Chen [19] correlation . . . . . . . . . . . . . . . . . . . . . . . . . . 139

A.3.5 Gungor and Winterton [52] correlation . . . . . . . . . . . . . . . . 140

A.3.6 Shah [130] correlation . . . . . . . . . . . . . . . . . . . . . . . . . . 140

A.3.7 Schrock and Grossman [129] correlation . . . . . . . . . . . . . . . . 141

A.3.8 Dembi et al. [30] correlation . . . . . . . . . . . . . . . . . . . . . . 141

A.3.9 Klimenko [84] correlation . . . . . . . . . . . . . . . . . . . . . . . . 141

A.3.10 Jung et al. [64] correlation . . . . . . . . . . . . . . . . . . . . . . . 142

A.4 Two phase flow: Mixture . . . . . . . . . . . . . . . . . . . . . . . . . . . . 142

A.4.1 Steiner [140] correlation . . . . . . . . . . . . . . . . . . . . . . . . 142

A.4.2 Kandlikar [71] correlation . . . . . . . . . . . . . . . . . . . . . . . 143

A.4.3 Bennett and Chen [8] correlation . . . . . . . . . . . . . . . . . . . 143

A.4.4 Palen [111] correlation . . . . . . . . . . . . . . . . . . . . . . . . . 143

A.4.5 Jung et al. [64] correlation . . . . . . . . . . . . . . . . . . . . . . . 144

B Pressure drop 145

B.1 Single phase . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 145

B.2 Two phase . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 145

B.2.1 Friedel [42] model . . . . . . . . . . . . . . . . . . . . . . . . . . . 147

B.2.2 Lockhart and Martinelli [91] model . . . . . . . . . . . . . . . . . . 147

B.2.3 Chisholm [22] model . . . . . . . . . . . . . . . . . . . . . . . . . . 148

C Physical properties 149

C.1 Physical properties: Pure fluid . . . . . . . . . . . . . . . . . . . . . . . . . 149

C.1.1 Specific heat . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149

C.1.2 Vapor pressure . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149

C.1.3 Liquid viscosity . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 149

C.1.4 Vapor dynamic viscosity VDI-Warmeatlas [157] . . . . . . . . . . . 149

C.1.5 Dynamic viscosity of Fenghour et al. [40] . . . . . . . . . . . . . . . 151

C.1.6 Surface tension . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 152

C.1.7 Thermal conductivity for liquids . . . . . . . . . . . . . . . . . . . . 152

C.1.8 Thermal conductivity for gases . . . . . . . . . . . . . . . . . . . . 152

C.1.9 Specific enthalpy . . . . . . . . . . . . . . . . . . . . . . . . . . . . 153

C.2 Physical properties: Mixture . . . . . . . . . . . . . . . . . . . . . . . . . . 153

C.2.1 Liquid dynamic viscosity of mixtures . . . . . . . . . . . . . . . . . 153

C.2.2 Vapor dynamic viscosity of mixtures . . . . . . . . . . . . . . . . . 153

C.2.3 Liquid thermal conductivity of mixtures . . . . . . . . . . . . . . . 154


Table of contents 7

C.2.4 Vapor thermal conductivity of mixtures . . . . . . . . . . . . . . . . 154

C.2.5 Surface tension of mixtures . . . . . . . . . . . . . . . . . . . . . . 155

C.3 Software packages . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 155


8 1 Introduction

1 IntroductionHeat exchanger is an important and expensive item of equipment that is used almost inevery industry (oil and petrochemical, sugar, food, pharmaceutical and power industry).A better understanding of the basic principles of heat transfer and fluid flow and theirapplication to the design and operation of heat exchangers that you gain from this coursewill enable you to improve their efficiency and extend their life. You understand how to usethe applicable API, TEMA and ASME recommended practices, standards and codes forheat exchangers. This will enable you to communicate with the designers, manufacturersand bidders of heat exchangers. You will understand how to avoid fouling, corrosion andfailure and leak problems by your design. You will also be able to survey and troubleshootheat exchangers and assist in performing inspection, cleaning, and maintenance. You willbe exposed to recent development and future trend in heat exchangers.

The course includes worked examples to reinforce the key learning as well as a demon-stration of mechanical design and challenging problems encountered in the operation ofheat exchangers.

Objectives

• To learn the classification, code and standards (API, TEMA,...) and selection pro-cedure for heat exchangers.

• To review the thermal and mechanical design of heat exchangers.

• To learn the installation, operation and maintenance procedure for heat exchanger.

• To acquire information that will enable decisions to be made on the repair andrefurbishment of aging equipment as well as repair vs. replacement options.

• To learn techniques of failure elimination and appropriate maintenance and trou-bleshooting procedures.

• To delineate the factors that lead to overall economically advantageous decisions.

Who should attend: Project engineers, process engineers and plant engineers in the oil,chemical, sugar, power, and other industries who requires a wider and deeper appreciationof heat exchangers design, performance and operation. The detailed review of thermaland mechanical design is particularly useful to plant and maintenance engineers as wellas to those generally knowledgeable in the subject, but who require a refresher or up-date. Codes and standards are useful for project engineer to help him communicate withmanufacturers, designers and bidders of heat exchangers. Troubleshooting procedures areimportant for process engineers. Participants will be taken through an intensive primerof heat transfer principles as applicable to heat exchangers.

1.1 Programm outline1. 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, componentdesign, nozzle loadings, supports, lifting features, high pressure, low tempera-ture, specials designs


1.1 Programm outline 9

2. DAY II HEAT TRANSFER FUNDAMENTALS AND THERMAL DE-SIGN

• Heat transfer mechanisms: conduction and convection as related to heat ex-changers

• Temperature difference in heat exchanger:

– LMTD Method– ε-NTU Method– θ-Method

• Overall heat transfer coefficient

• Heat transfer coefficient and pressure drop for single phase and multiphase(evaporation and condensation)

• Resistances to fouling

• Illustration examples using the software CHEMCAD

3. DAY III MECHANICAL DESIGN OF HE

• Mechanical design: shells, channels and heads, tubesheets, bundles, tubes-tubesheet attachment

• Design strategy, design algorithm

• Heat exchanger:

– Selection procedure– Specification sheet– Bid evaluation

• Worked example (USING CHEMCAD)

4. DAY IV Storage, Installation, Operation, Maintenance

• Storage

• Installation procedure

• Operation

• start up

• shut down

• Maintenance

• Cleaning

• Repair

– Plug– Sleeving– Expansion

• Replacement

– Retubing– Rebundling– Replacement (new unit)

5. DAY V Troubleshooting

• Heat exchangers’ problem


10 1 Introduction

– Fouling: causes, mechanisms, design considerations and exchanger selec-tion, remedies, cleaning

– Leakage: Location (tube sheet, tube failure), causes (differential thermalexpansion, flow-induced vibration),

– Corrosion: Type, causes, material of construction, fabrication– Vibration: causes (velocity), design procedure to avoid vibration including

baffle selection, rod baffles, impingement baffles

• Past incidents failure.

• Examples of common problems encountered in heat exchangers (low rate, un-controlled outlet temperature, failure of tubes near the inlet nozzles)

Achieve the learning outcomes to:

Understand the principles of heat transfer and fluid flow, application of industry prac-tices and a substantial amount of supporting data needed for design, performance andoperation of modern heat exchangers.

Gain insight not only into shell and tube heat exchangers but also heat transfer funda-mentals as applied to heat exchangers, the types of heat exchangers and their application,and recent advance in heat exchanger technologies

Become familiar with the practical aspects and receive tips on shell and tube heatexchanger thermal design and rating: mechanical design and rating using the applicableAPI, 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 determiningand correcting them

Daily Schedule: 8:00 Registration and Coffee (1st day only) 8:30 Session begins 4:30Adjournment

There will be a forty-minute lunch break each day in addition to refreshment and net-working break of 20 minutes during each morning and afternoon session.

1.2 InstructorFaculty: Ali. Rabah, BSc. MSc., PhD., MSES., Assistant professor, De-partment of Chemical Engineering University of Khartoum

Dr. Rabah holds a BSc. degree (Chemical Engineering) from the University of Khartoum,MSc. degree from university of Nairobi, Kenya, and PhD. degree from University ofHannover, Germany. He has a wide professional experience in teaching heat and masstransfer and engineering thermodynamics to BSc and MSc Chemical, Mechanical andPetroleum 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 andboilers. He designed, for example, a 5 ton/hr (10 bar) fired tube boiler. His design isunder fabrication.

Dr. Rabah has designed and manufactured double pipe heat exchangers for educationproposes to a number of chemical engineering departments country-wide e.g. Universityof Nileen.

Dr. Rabah assumed engineering design positions with responsibilities covering design,construction and inspection of heat transfer equipments. The design projects are spon-sored by the federal ministry of research and technology and the University of Khartoumconsultancy cooperation.


1.2 Instructor 11

Dr. Rabah is a member of the Sudan Engineering Society (SES) and serving as a memberof editorial board of SES Journal. He is a reviewer to a number of world wide soft-ware packages for chemical engineering simulations and the prediction of thermodynamicproperties.

Dr. Rabah has a number of publications in field of heat transfer and thermodynamics.


12 2 Classification of heat exchangers

2 Classification of heat exchangersThe word exchanger really applies to all types of equipment in which heat is exchanged butis often used specially to denote equipment in which heat is exchanged between two processstreams. Exchangers in which a process fluid is heated or cooled by a plant service streamare referred to as heatsers and coolers. If the process stream is vaporized the exchanger iscalled a vaporizer if the the stream is essentially completely vaporized: called a reboiledif 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 firedexchanger is used for exchangers heated by combustion gases, such as boiler. In heatexchanger the heat transfer between the fluid takes place through a separating wall. Thewall may a solid wall or interface. Heat exchangers are used in

• Oil and petrochemical Industry (upstream and down stream)

• Sugar industry

• Power generation industry

• Air-cooling and refrigeration industry

These heat exchanger may be classified according to:

• Transfer process

1. Direct contact

2. indirect contact

(a) Direct transfer type(b) Storage type(c) Fluidized bed

• Surface compactness

1. Compact (surface area density ≥ 700m2/m3)

2. non-compact (surface area density < 700m2/m3)

• Construction

1. Tubular

(a) Double pipe(b) Shell and tube(c) Spiral tube

2. Plate

(a) Gasketed(b) Spiral plate(c) Welded plate

3. Extended surface

(a) Plate fin(b) Tube fin

4. Regenerative

(a) Rotoryi. Disc-type


13

ii. Drum-type(b) Fixed-matrix

• Flow arrangement

1. Single pass

(a) Parallel flow(b) Counter flow(c) Cross flow

2. Multipass

(a) Extended surface H.E.i. Cross counter flowii. 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 flowiii. 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 thecooling or heating of a liquid or gas) or two-phase (such as condensing or vaporizing).Since there are two sides to an STHE, this can lead to several combinations of ser-vices. Broadly, services can be classified as follows: single-phase (both shellside andtubeside); 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 isusually used:

– Heat exchanger: both sides singlephase and process streams (that is, not autility).

– Cooler: one stream a process fluid and the other cooling water or air. Dirtywater can be used as the cooling medium. The top of the cooler is open to theatmosphere for access to tubes. These can be cleaned without shutting downthe 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 steamor hot oil.

– Condenser: one stream a condensing vapor and the other cooling water or air.



– Chiller: one stream a process fluid being condensed at sub-atmospheric tem-peratures and the other a boiling refrigerant or process stream. By cooling thefalling film to its freezing point, these exchangers convert a variety of chemicalsto the solid phase. The most common application is the production of sized iceand paradichlorobenzene. Selective freezing is used for isolating isomers. Bymelting the solid material and refreezing in several stages, a higher degree ofpurity of product can be obtained.

– Reboiler: one stream a bottoms stream from a distillation column and theother a hot utility (steam or hot oil) or a process stream.

– Evaporators:These are used extensively for the concentration of ammoniumnitrate, urea, and other chemicals sensitive to heat when minimum contacttime is desirable. Air is sometimes introduced in the tubes to lower the partialpressure of liquids whose boiling points are high. These evaporators are builtfor pressure or vacuum and with top or bottom vapor removal.

– Absorbers: These have a two-phase flow system. The absorbing medium isput in film flow during its fall downward on the tubes as it is cooled by a coolingmedium outside the tubes. The film absorbs the gas which is introduced intothe tubes. This operation can be cocurrent or countercurrent.

– Falling-Film Exchangers: Falling-film shell-and-tube heat exchangers havebeen 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 verticaltubes. Distributors or slotted tubes put the liquid in film flow in the insidesurface of the tubes, and the film adheres to the tube surface while fallingto the bottom of the tubes. The film can be cooled, heated, evaporated, orfrozen by means of the proper heat-transfer medium outside the tubes. Tubedistributors have been developed for a wide range of applications. Fixed tubesheets, with or without expansion joints, and outside-packed-head designs areused. Principal advantages are high rate of heat transfer, no internal pressuredrop, short time of contact (very important for heat-sensitive materials), easyaccessibility to tubes for cleaning, and, in some cases, prevention of leakagefrom one side to another. These falling-film exchangers are used in variousservices as described in the following paragraphs.

Among these classifications the classification by construction is the most widely used one.

2.1 Classification by constructionThe 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 isfurther 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 exchangerThis is usually consists of concentric pipes. One fluid flow in the inner pipe and the otherfluid flow in the annulus between pipes. The two fluid may flow concurrent (parallel) orin 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

Flo

wm

eter

Bypass

pump

Tee 2"x1/2"

Union 2"

Galv. pipe 2"

Cu pipe 3/4"

Tee 3/4"x1/2"

Elbew 3/4"

Flanged Gland 2"

Part B

Double Pipe Heat ExchangerScale: None Sheet No.1 Date: 08.12.2003

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

Part A

Specification Sheet

Item Qty Item Qty

Tee 2"x3/4" 6 Tee 3/4"x1/2" 14

Union 2" 6 Cu Bush 1/2" 8

Valve 3/4" 4 Elbew 3/4" 10

Galv. pipe 2"x3ft 3 Cu pipe 3/4"x4ft 3

Galv. pipe 3/4"x1ft Selector

(Threaded) 24 (20 Channel) 1

Cu Flange 2" 8 Flow meter 3/4" 2

Pump 0-40 l/min 2 Union 3/4" 30

Amplifier 1 Microvoltmeter 1

Thermocouples Elbew 1/2" 4

(NiCr-Ni) 10 Union 1/2" 8

Val

ve3/

4"

Galv. pipe

Threaded 3/4"

Bypass

Figure 2.2. Double pipe heat exchanger (Counter current)

Double pipe heat exchanger is perhaps the simplest of all heat exchanger types. Theadvantages of this type are:



i Easily by disassembly, no cleaning problem

ii Suitable for high pressure fluid, (the pressure containment in the small diameter pipeor tubing is a less costly method compared to a large diameter shell.)

Limitation: The double pipe heat exchanger is generally used for the application wherethe total heat transfer surface area required is less than or equal to 20 m2 (215 ft2) becauseit is expensive on a cost per square meter (foot) basis.

2.3 Spiral tube heat exchangerSpiral 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 straighttube . In addition, considerable amount of surface area can be accommodated in a givenspace 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 exchangerShell and tube heat exchanger is built of round tubes mounted in a cylindrical shell withthe tube axis parallel to that of the shell. One fluid flow inside the tube, the other flowacross and along the tubes. The major components of the shell and tube heat exchangerare 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 bothends to tubesheets welded to the shell. The construction may have removable channelcovers , bonnet-type channel covers , or integral tubesheets. The principal advantage ofthe 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 jointis required.

Figure 2.5. Fixed-tubesheet heat exchanger.

Other advantages are that the tubes can be cleaned mechanically after removal of thechannel cover or bonnet, and that leakage of the shellside fluid is minimized since thereare no flanged joints.

A disadvantage of this design is that since the bundle is fixed to the shell and cannot beremoved, the outsides of the tubes cannot be cleaned mechanically. Thus, its applicationis limited to clean services on the shellside. However, if a satisfactory chemical clean-ing program can be employed, fixed-tubesheet construction may be selected for foulingservices on the shellside.

In the event of a large differential temperature between the tubes and the shell, thetubesheets will be unable to absorb the differential stress, thereby making it necessary toincorporate an expansion joint. This takes away the advantage of low cost to a significantextent.

2.4.2 U-tube

As the name implies, the tubes of a U-tube heat exchanger (Figure 2.6) are bent inthe 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 thebending of the tubes and the somewhat larger shell diameter (due to the minimum U-bendradius), making the cost of a U-tube heat exchanger comparable to that of a fixedtubesheetexchanger.



The advantage of a U-tube heat exchanger is that because one end is free, the bundlecan expand or contract in response to stress differentials. In addition, the outsides of thetubes 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 becleaned 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 insidetubes.

Figure 2.6. U-tube heat exchanger.

2.4.3 Floating head

The floating-head heat exchanger is the most versatile type of STHE, and also the costliest.In this design, one tubesheet is fixed relative to the shell, and the other is free to ”float”within the shell. This permits free expansion of the tube bundle, as well as cleaningof both the insides and outsides of the tubes. Thus, floating-head SHTEs can be usedfor services where both the shellside and the tubeside fluids are dirty-making this thestandard construction type used in dirty services, such as in petroleum refineries.

There are various types of floating- head construction. The two most common are thepull-through with backing device and pullthrough without backing service designs. Thedesign (Figure 2.7) with backing service is the most common configuration in the chemicalprocess industries (CPI). The floating-head cover is secured against the floating tubesheetby bolting it to an ingenious split backing ring. This floating-head closure is locatedbeyond the end of the shell and contained by a shell cover of a larger diameter. Todismantle 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 thestationary 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 theshell diameter is larger than the floating-head flange. The floatinghead cover is bolteddirectly to the floating tubesheet so that a split backing ring is not required. The advan-tage of this construction is that the tube bundle may be removed from the shell withoutremoving either the shell or the floatinghead cover, thus reducing maintenance time. Thisdesign is particularly suited to kettle reboilers having a dirty heating medium where U-tubes cannot be employed. Due to the enlarged shell, this construction has the highestcost of all exchanger types.

Figure 2.8. Floating head without packing service.

2.5 Plate heat exchangersThese exchangers are generally built of thin plates. The plate are either smooth or havesome form of corrugations and they are either flat or wound in exchanger. Generally



theses exchanger cannot accomodate high pressure/temperature differential relative thetubular 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 alloymaterial channel plates, bounded by elastomeric gaskets are hung off and guided by lon-gitudinal carrying bars, then compressed by large-diameter tightening bolts between twopressure retaining frame plates (cover plates).

Figure 2.9. Plate heat exchanger

The frame and channel plates have portholes which allow the process fluids to enter alter-nating flow passages (the space between two adjacent-channel plates) Fig.2.10. Gasketsaround the periphery of the channel plate prevent leakage to the atmosphere and also pre-vent process fluids from coming in contact with the frame plates. No inter fluid leakageis 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 morechannel plates (i.e. surface area). In fact, if a known future capacity is available duringfabrication stages, a longer carrying bar could be installed, and later, increasing thesurface area would be easily handled. When the expansion is needed, simply untightenthe carrying bolts, pull back the frame plate, add the additional channel plates, andtighten the frame plate.



Figure 2.10. Plate heat exchanger flow configuration

Applications: Most PHE applications are liquid-liquid services but there are numeroussteam heater and evaporator uses from their heritage in the food industry. Industrial userstypically have chevron style channel plates while some food applications are washboardstyle.

Fine particulate slurries in concentrations up to 70 percent by weight are possible withstandard channel spacings. Wide-gap units are used with larger particle sizes. Typicalparticle 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 theability to sanitize the entire heat transfer surface easily were a major benefit in the foodand 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 thefluid split on the floor rather than mixing with other fluid

• Heat transfer coefficient is larger and hence small heat transfer area is required thanSTHE

• The space required is less than that for STHE for the same duty

• Less fouling due to high turbulent flow



Figure 2.11. Plate and frame of a plate heat exchanger

• Very close temperature approach can be obtained

• low hold up volume

• LMTD is fully utilized

• More economical when material cost are high

Disadvantages: -

• Low pressure <30 bar (plate deformation)

• Working temperature of < (500 F) [250 oC] (maximum gasket temperature) seetable 2.1.

Table 2.1. Plate Heat Exchanger Gasket MaterialsMaterial Common name Temperature limit (F)Styrene-Butadiene Buna-S 185Neoprene Neoprene 250Acrylonitrile- Butadiene Buna-N 275Ethylene/Propylene EPDM 300Fluorocarbon Viton 300Resin-Cured Butyl Resin-Cured Butyl 300Compressed Asbestos Compressed Asbestos 500

2.5.2 Welded- and Brazed-Plate exchanger (W. PHE and BHE)

To overcome the gasket limitations, PHE manufacturers have developed welded-plateexchangers. There are numerous approaches to this solution: weld plate pairs togetherwith the other fluid-side conventionally gasketed, weld up both sides but use a horizonal



stacking of plates method of assembly, entirely braze the plates together with copper ornickel brazing, diffusion bond then pressure form plates and bond etched, passage platesFig. 2.12 and Fig. 2.13.

Typical applications include district heating where the low cost and minimal maintenancehave made this type of heat exchanger especially attractive.

Figure 2.12. Welded or blazed plate heat exchanger

Figure 2.13. Fin-Plate heat exchanger

Most methods of welded-plate manufacturing do not allow for inspection of the heat-transfer surface, mechanical cleaning of that surface, and have limited ability to repairor 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 itsvirtues and not on its initial cost. SPHEs offer high reliability and on-line performance inmany 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.2.14and Fig.2.15. This forms two passages. Passages are sealed off on one end of the SHE bywelding a bar to the plates; hot and cold fluid passages are sealed off on opposite ends ofthe SHE. A single rectangular flow passage is now formed for each fluid, producing veryhigh shear rates compared to tubular designs. Removable covers are provided on eachend 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 entrainany solids and create high turbulence creating a self-cleaning flow passage. There areno thermal-expansion problems in spirals. Since the center of the unit is not fixed, itcan 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 befabricated out of the higher alloy. SHEs can be fabricated out of any material that can becold-worked and welded. The channel spacings can be different on each side to match theflow rates and pressure drops of the process design. The spacer studs are also adjusted intheir pitch to match the fluid characteristics. As the coiled plate spirals outward, the platethickness 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 totubing of conventional exchangers.

a) Spiral flow in both channels b) Flow are both spiral and axial

Figure 2.15. Spiral Plate heat exchanger

Applications: The most common applications that fit SHE are slurries. The rectan-gular channel provides high shear and turbulence to sweep the surface clear of blockageand causes no distribution problems associated with other exchanger types. A localizedrestriction 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. Thisprovides some degree of corrosion protection as well, but this is not guaranteed due topinholes 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 andcooling services and can accommodate temperature crosses such as lean/rich servicesin one unit. The removable covers on each end allow access to one side at a time toperform maintenance on that fluid side. Never remove a cover with one side underpressure 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 spiralflow and the other side is in cross flow. These SHEs provide very stable designsfor vacuum condensing and reboiling services. A SHE can be fitted with specialmounting connections for reflux-type ventcondenser applications. The verticallymounted 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 spiralflow and part is in cross flow. This SHE can condense and subcool in a singleunit. 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 previouslyflowed there) off the surface. Phosphoric acid coolers use pond water for coolingand both sides foul; water, as you expect, and phosphoric acid deposit crystals. Byreversing the flow sides, the water dissolves the acid crystals and the acid clears upthe 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 notperformed well.

2.6 Extended surfaceThe tubular and plate exchangers described previously are all prime surface heat exchang-ers. The design thermal effectiveness is usually 60 % and below and the heat transfer areadensity is usually less than 300 m2m3. In many application an effectiveness of up to 90% is essential and the box volume and mass are limited so that a much more compactsurface is mandated. Usually either a gas or a liquid having a low heat transfer coefficientis 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 ina reasonably shaped configuration.

The surface area may be increased by the fins. The flow area is increased by the use ofthin gauge material and sizing the core property. There are two most common types ofextended 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 (refereedto as parting plates or parting sheets) or formed tubes as shown in fig. 2.16(left). Whilethe plates separate the two fluid streams, the fins form the individual flow passages. Finsare used on both sides in a gas-gas heat exchanger. In gas-liquid applications fins areused 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 generallyused. Fins are generally used on the outside and also used inside the tubes in someapplications. 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 standardsThe objective of codes and standards are best described by ASME: The objectives ofcode rules and standards (apart from fixing dimensional values) is to achieve minimumrequirements for safe construction, in other words, to provide public protection by definingthose materials, design, fabrication and inspection requirements; whose omission mayradically increase operating hazards.... Experience with code rules has demonstrated thatthe probability of disastrous failure can be reduced to the extremely low level necessary toprotect life and property by suitable minimum requirements and safety factors. Obviously,it is impossible for general rules to anticipate other than conventional service,.... Suitableprecautions are therefore entirely the responsibility of the design engineer guided by theneeds 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 andtube 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 afew.

In this work, being most widely used one, the TEMA standard is presented.

3.1 TEMA DesignationsIn order to understand the design and operation of the shell and tube heat exchanger, itis important to know the nomenclature and terminology used to describe them and thevarious parts that go to their construction. Only then we can understand the design andreports given by the researchers, designers, manufacturer and users.

It is essential for the designer to have a good working knowledge of the mechanical featuresof STHEs and how they influence thermal design. The principal components of an STHEare:

• shell;

• shell cover;

• tubes;

• channel;

• channel cover;

• tubesheet;

• baffles; and

• nozzles.

Other components include tie-rods and spacers, pass partition plates, impingement plate,longitudinal baffle, sealing strips, supports, and foundation. Table 3.1 shows the nomen-clature used for different parts of shell and tube exchanger in accordance with TEMAstandards; the numbers refer to the feature shown in Fig. 3.2 to Fig. 3.8.


3.1 TEMA Designations 29

Table 3.1. TEAM notationsIndex Notation Index Notation

1 stationary head- channel 20 slip on backing flange2 stationary head- bonnet 21 floating head cover-external3 stationary head flange-chennel or bonnet 22 floating tube sheet skirt4 channel cover 23 packing box5 stationary head - nozzle 24 packing6 stationary tube sheet 25 packing gland7 tubes 26 latern ring8 shell 27 tie rods and spacers9 shell cover 28 traverse baffle or support plate10 shell flange-stationary head end 29 impingement plate11 shell flange-rear head end 30 longitudinal baffle12 shell nozzle 31 pass partition13 shell cover flange 32 vent connection14 expansion joint 33 drain connection15 floating tube sheet 34 instrument connection16 floating head cover 35 support saddle17 floating head flange 36 lifting lug18 floating head backing device 37 support bracket19 split shear ring 38 weir

39 liquid level connection

Because of the number of variations in mechanical designs for front and rear heads andshells, 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 variousconstruction possibilities. TEMA has classified the front head channel and bonnet types asgiven the letters (A,B,C,N,D) and the shell is classified according to the nozzles locationsfor the inlet and outlet. There are type of shell configuration ( E,F,G,H,J,K,X). Similarlythe 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 standsfor the front head, the second letter for the shell type and the third letter for the rear headtype. For example a BFL exchanger has a bonnet cover, two-shell pass with longitudinalbaffles 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 arerounded 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 channelcover (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 stationaryfront 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 coverfront head (A), two-pass shell (F) and fixed tube sheet bonnet-type rear head (M)with 331/8 in (840 mm) inside diameter and 8ft (2440 mm) tube length.



Figure 3.1. TEMA-type designations for shell-and-tube heat exchangers. (Standards of Tubu-lar Exchanger Manufacturers Association, 6th ed., 1978.)

In the above illustration the term single pass and two pass shell have been used. Thismean that the shell side fluid travels only one through the shell (single pass) or twice (twopass shell). Two pass shell mean that the fluid enters at one end, travel to other end andback to the end where it entered (making U-turn). Similarly there are multiple pases. Tobe remembered is that the number of tube passes is equal to or greater than the numberof shell passes. Generally the multi shell and tube passes are usually designated by twonumerals separated by a hyphen, with the first numeral indication the number of shellpass and the other stands for the tube passes. For example a one-shell pass and two tubepass AEL exchanger will be written as 1-2 AEL. To be remembered is that this not anTEMA 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 forindicating the number of shell and tube passes. Another identification of the shell andtube heat exchanger is the number of shell passes. 1 shell pass, 2 shell pass, etc. This isnot a TEMA standardization. The tube passes can be equal to or greater than the shellpass.


32 3 Code and standardsTab

le3.2.

Featuresof

TE

MA

Shell-and-Tube-T

ypeE

xchangers.

Packed

InternalO

utsideT

ypeFixed

lantern-ringfloating

headpacked

Pull-through

ofdesign

tubesheet

U-tube

floatinghead

(splitbacking

ring)floating

headfloating

headT

.E.M

.A.rear

-headtype

Lor

Mor

NU

WS

PT

Relative

costincreases

fromA

(leastexpensive)

throughE

(most

expensive)B

AC

ED

EP

rovisionfor

differentialexpansion

Expansion

Individualtubes

Floating

headFloating

headFloating

headFloating

headjoint

infree

toexpand

shellR

emovable

bundleN

oY

esY

esY

esY

esY

esR

eplacement

bundlepossible

No

Yes

Yes

Yes

Yes

Yes

Individualtubes

replaceableY

esO

nlythose

inY

esY

esY

esY

esoutside

rowTube

cleaningby

chemicals

insideand

outsideY

esY

esY

esY

esY

esY

esInterior

tubecleaning

mechanically

Yes

Specialtools

requiredY

esY

esY

esY

esE

xteriortube

cleaningm

echanically:Triangular

pitchN

oN

oN

oN

oN

oN

oSquare

pitchN

oY

esY

esY

esY

esY

esH

ydraulic-jetcleaning:Tube

interiorY

esSpecial

toolsrequired

Yes

Yes

Yes

Yes

Tube

exteriorN

oY

esY

esY

esY

esY

esD

oubletube

sheetfeasible

Yes

Yes

No

No

Yes

No

Num

berof

tubepasses

No

practicalA

nyeven

Lim

itedto

oneN

opractical

No

practicalN

opractical

limitations

number

possibleor

two

passeslim

itationslim

itationslim

itationsInternal

gasketselim

inatedY

esY

esY

esN

oY

esN

o


3.2 Classification by construction STHE 33

3.2 Classification by construction STHEFig. 3.2 to Fig. 3.8 show details of the construction of the TEMA types of shell-and-tubeheat 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, andthe frequency of use has been increasing in recent years. The tube sheets are weldedto the shell. Usually these extend beyond the shell and serve as flanges to which thetube-side headers are bolted. This construction requires that the shell and tube-sheetmaterials be weldable to each other. When such welding is not possible, a blind-gaskettype of construction is utilized. The blind gasket is not accessible for maintenance orreplacement once the unit has been constructed. This construction is used for steamsurface condensers, which operate under vacuum.

Figure 3.2. Heat-exchanger-component nomenclature. Fixed tube heat sheet shell and tubeheat 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. 3.1for type C and N heads. This type of construction is less costly than types B and M orA and L and still offers the advantage that tubes may be examined and replaced withoutdisturbing the tube-side piping connections. There is no limitation on the number oftube-side passes. Shell-side passes can be one or more, although shells with more thantwo 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 necessaryfor fabrication. Between the inside of the shell and the baffles some clearance must beprovided so that baffles can slide into the shell. Fabrication tolerances then require someadditional clearance between the outside of the baffles and the outermost tubes. The edgedistance between the outer tube limit (OTL) and the baffle diameter must be sufficientto prevent vibration of the tubes from breaking through the baffle holes. The outermosttube 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- through24-in) pipe shells, and slightly less for smaller-diameter pipe shells.



Tubes can be replaced. Tube-side headers, channel covers, gaskets, etc., are accessible formaintenance and replacement. Neither the shell-side baffle structure nor the blind gasketis 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 theappropriate holes in the tube sheets.

Differential expansion between the shell and the tubes can develop because of differencesin length caused by thermal expansion. Various types of expansion joints are used toeliminate excessive stresses caused by expansion. The need for an expansion joint is afunction of both the amount of differential expansion and the cycling conditions to beexpected 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 platescan flex to make some allowance for differential expansion. This design is generallyused for vacuum service and gauge pressures below 103 kPa (15 lbf/in2). All weldsare subject to severe stress during differential expansion.

b Flanged-only heads. The flat plates are flanged (or curved). The diameter of theseheads is generally 203 mm (8 in) or more greater than the shell diameter. Thewelded 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 thecurved shape.

c Flared shell or pipe segments. The shell may be flared to connect with a pipesection, 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 canbe used. These are welded together or connected by a ring. This type of joint issimilar 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 concentricreverse flue holes. These heads are relatively expensive because of the cost of thefluing operation. The curved shape of the heads reduces the amount of stress at thewelds to the shell and also connecting the heads.

f Toroidal. The toroidal joint has a mathematically predictable smooth stress pat-tern of low magnitude, with maximum stresses at sidewalls of the corrugation andminimum stresses at top and bottom. The foregoing designs were discussed as ringexpansion joints by Kopp and Sayre, Expansion Joints for Heat Exchangers (ASMEMisc. Pap., vol. 6, no. 211). All are statically indeterminate but are subjectedto analysis by introducing various simplifying assumptions. Some joints in currentindustrial use are of lighter wall construction than is indicated by the method ofthis paper.

g Bellows. Thin-wall bellows joints are produced by various manufacturers. These aredesigned for differential expansion and are tested for axial and transverse movementas well as for cyclical life. Bellows may be of stainless steel, nickel alloys, or copper.(Aluminum, Monel, phosphor bronze, and titanium bellows have been manufac-tured.) Welding nipples of the same composition as the heat-exchanger shell aregenerally furnished. The bellows may be hydraulically formed from a single pieceof metal or may consist of welded pieces. External insulation covers of carbon steelare often provided to protect the light-gauge bellows from damage. The cover alsoprevents insulation from interfering with movement of the bellows (see h).

h Toroidal bellows. For high-pressure service the bellows type of joint has been modi-fied so that movement is taken up by thin-wall small-diameter bellows of a toroidalshape. Thickness of parts under high pressure is reduced considerably (see f ).

Improper handling during manufacture, transit, installation, or maintenance of the heatexchanger equipped with the thin-wallbellows type or toroidal type of expansion joint candamage the joint. In larger units these light-wall joints are particularly susceptible todamage, and some designers prefer the use of the heavier walls of formed heads.

Chemical-plant exchangers requiring expansion joints most commonly have used theflanged-and-flued-head type. There is a trend toward more common use of the light-wall-bellows type.

3.2.2 U-Tube Heat Exchanger

Fig. 3.4 shows U-tube heat exchanger Type CFU. The tube bundle consists of a stationarytube sheet, U tubes (or hairpin tubes), baffles or support plates, and appropriate tie rodsand spacers. The tube bundle can be removed from the heat-exchanger shell. A tube-sideheader (stationary head) and a shell with integral shell cover, which is welded to theshell, are provided. Each tube is free to expand or contract without any limitation beingplaced upon it by the other tubes. The U-tube bundle has the advantage of providingminimum clearance between the outer tube limit and the inside of the shell for any ofthe removable-tube-bundle constructions. Clearances are of the same magnitude as forfixed-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 veryshort 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-pressureconstruction this feature becomes of considerable importance in reducing both initial andmaintenance costs. The use of U-tube construction has increased significantly with thedevelopment of hydraulic tube cleaners, which can remove fouling residues from both thestraight and the U-bend portions of the tubes. Rods and conventional mechanical tubecleaners 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. Hydraulicjetting with water forced through spray nozzles at high pressure for cleaning tube interiorsand 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 designis often used with outdoor storage tanks for heavy fuel oils, tar, molasses, and similarfluids whose viscosity must be lowered to permit easy pumping. Uusally the tube-sideheating medium is steam. One end of the heater shell is open, and the liquid being heatedpasses across the outside of the tubes. Pumping costs can be reduced without heating theentire 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 thefouling 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 shellsections for vapor-liquid separation (Fig.3.6). The U-tube bundle replaces the floating-heat bundle of Fig. 3.4.

Figure 3.6. Kettle reboiler

The U-tube exchanger with copper tubes, cast-iron header, and other parts of carbonsteel 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 frequentlyused substitute materials. These standard exchangers are available from a number ofmanufacturers 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 bundlefor cleaning, several floating rear head designs have been established. The simplest is a



Internal floating head (pullthrough design) Fig3.9 design which allows the tube bundle tobe pulled entirely through the shell for service or replacement. In order to accommodatethe rear head bolt circle, tubes must be removed resulting in a less efficient use of shellsize. In addition, the missing tubes result in larger annular spaces and can contribute toreduced flow across the effective tube surface, resulting in reduced thermal performance.

Some designs include sealing strips installed in the shell to help block the bypass steam.Another floating head design that partially addresses the above disadvantages is a split-ring floating head. Here the floating head bonnet is bolted to a split backing ring insteadof the tube sheet. This eliminates the bolt circle diameter and allows a full complementof tubes to fill the shell. This construction is more expensive than a common pull throughdesign, but is in wide use in petrochemical applications. For applications with highpressures 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 boxdesigns offer less positive sealing against leakage to the atmosphere than the pull thoughor split ring designs, but can be configured for single tube pass duty. More details aboutthe various types of floating head shell and tube heat exchanger is given the followingsections

Packed-Lantern-Ring Exchanger: (Fig. 3.7 ) This construction is the least costlyof the straight-tube removable bundle types. The shell- and tube-side fluids are eachcontained by separate rings of packing separated by a lantern ring and are installed at thefloating tube sheet. The lantern ring is provided with weep holes. Any leakage passingthe packing goes through the weep holes and then drops to the ground. Leakage at thepacking will not result in mixing within the exchanger of the two fluids. The width of thefloating tube sheet must be great enough to allow for the packings, the lantern ring, anddifferential expansion. Sometimes a small skirt is attached to a thin tube sheet to providethe required bearing surface for packings and lantern ring. The clearance between theouter tube limit and the inside of the shell is slightly larger than that for fixed-tube-sheetand U-tube exchangers.

The use of a floating-tube-sheet skirt increases this clearance. Without the skirt theclearance must make allowance for tubehole distortion during tube rolling near the outsideedge 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 below191C (375F) and to the mild services of water, steam, air, lubricating oil, etc. Designgauge pressure does not exceed 2068 kPa (300 lbf/in2) for pipe shell exchangers and islimited to 1034 kPa (150 lbf/in2) for 610- to 1067-mm- (24- to 42-in-) diameter shells.



Figure 3.7. Heat-exchanger-component nomenclature. Exchanger with packed floating tubesheet and lantern ring. Type AJW. External floating head design. (Standard of Tubular Ex-changer Manufacturers Association, 6th ed., 1978.)

Outside-Packed Floating-Head Exchanger: (Fig. 3.8) The shell-side fluid is con-tained by rings of packing, which are compressed within a stuffing box by a packingfollower ring. This construction was frequently used in the chemical industry, but inrecent years usage has decreased. The removable-bundle construction accommodates dif-ferential expansion between shell and tubes and is used for shell-side service up to 4137kPa gauge pressure (600 lbf/in2) at 316C (600F).

Figure 3.8. Heat-exchanger-component nomenclature. Outside-packed floating-head ex-changer. Type AEP. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)

There are no limitations upon the number of tube-side passes or upon the tube-sidedesign pressure and temperature. The outside-packed floating-head exchanger was themost 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 machinefinish. A split shear ring is inserted into a groove in the floating-tube-sheet skirt. A slip-on backing flange, which in service is held in place by the shear ring, bolts to the externalfloating- head cover. The floating-head cover is usually a circular disk. With an oddnumber 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 channelbarrel (similar to Fig. 11-36f ) bolted between floating-head cover and floating-tube-sheetskirt. The outer tube limit approaches the inside of the skirt but is farther removed fromthe 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 moderatediameterplate shells.

Internal Floating-Head Exchanger: (Fig. 3.9) The internal floating-head designis used extensively in petroleum-refinery service, but in recent years there has been adecline in usage. The tube bundle is removable, and the floating tube sheet moves (orfloats) to accommodate differential expansion between shell and tubes. The outer tubelimit 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 plateshells. A split backing ring and bolting usually hold the floating-head cover at the floatingtube sheet. These are located beyond the end of the shell and within the larger-diametershell cover. Shell cover, split backing ring, and floating-head cover must be removed beforethe tube bundle can pass through the exchanger shell. With an even number of tube-sidepasses the floating-head cover serves as return cover for the tube-side fluid. With an oddnumber of passes a nozzle pipe must extend from the floating-head cover through the shellcover. Provision for both differential expansion and tube-bundle removal must be made.

Figure 3.9. Heat-exchanger-component nomenclature. Internal floating head (pull- throughdesign). Type AES. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)

Figure 3.10. Heat-exchanger-component nomenclature. Exchanger with packed floating tubesheet 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 thatof 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 fromthe shell without removing either shell cover or floating-head cover. This feature reducesmaintenance time during inspection and repair.

The large clearance between the tubes and the shell must provide for both the gasketand the bolting at the floating-head cover. This clearance is about 2 to 2.5 times thatrequired by the split-ring design. Sealing strips or dummy tubes are often installed toreduce 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, othershells offer distinct advantages. For example, the TEMA-F shell design providesfor a longitudinal flow plate to be installed inside the tube bundle assembly. Thisplate causes the shell fluid to travel down one half of the tube bundle, then downthe other half, in effect producing a counter-current flow pattern which is best forheat transfer. This type of construction can be specified where a close approachtemperature is required and when the flow rate permits the use of one half of theshell at a time. In heat recovery applications, or where the application calls forincreased thermal length to achieve effective overall heat transfer, shells can beinstalled with the flows in series. Up to six shorter shells in series is common andresults in counter-current flow close to performance as if one long shell in a singlepass design were used.

• TEMA G and H shell designs are most suitable for phase change applications wherethe bypass around the longitudinal plate and counter-current flow is less impor-tant than even flow distribution. In this type of shell, the longitudinal plate offersbetter flow distribution in vapor streams and helps to flush out non-condensable.They are frequently specified for use in horizontal thermosiphon reboilers and totalcondensers.

• TEMA J Shells are typically specified for phase change duties where significantlyreduced shell side pressure drops are required. They are commonly used in stackedsets with the single nozzles used as the inlet and outlet. A special type of J-shellis used for flooded evaporation of shell side fluids. A separate vapor disengagementvessel without tubes is installed above the main J shell with the vapor outlet at thetop of this vessel. The



• TEMA K shell, also termed a kettle reboiler, is specified when the shell side streamwill undergo vaporization. The liquid level of a K shell design should just cover thetube bundle, which fills the smaller diameter end of the shell. This liquid level iscontrolled by the liquid flowing over a weir at the far end of the entrance nozzle. Theexpanded shell area serves to facilitate vapor disengagement for boiling liquid in thebottom of the shell. To insure against excessive liquid carry-though with the vaporstream, a separate vessel as described above is specified. Liquid carry-through canalso be minimized by installing a mesh demister at the vapor exit nozzle. U-bundlesare typically used with K shell designs. K shells are expensive for high pressurevaporization due to shell diameter and the required wall thickness.

• The TEMA X shell, or crossflow shell is most commonly used in vapor condensingapplications, though it can also be used effectively in low pressure gas cooling orheating. It produces a very low shell side pressure drop, and is therefore mostsuitable for vacuum service condensing. In order to assure adequate distributionof vapors, X-shell designs typically feature an area free of tubes along the top ofthe exchanger. It is also typical to design X shell condensers with a flow area atthe 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-shellconstructions.

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 inorder to examine the tube ends. The fixed-tubesheet exchanger of Fig. 3.1b hasbonnets at both ends of the shell.

• The channel (Fig. 3.1A) has a removable channel cover. The tube ends can beexamined by removing this cover without disturbing the piping connections to thechannel 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. Thisdesign is comparable in cost with the bonnet but has the advantages of permittingaccess to the tubes without disturbing the piping connections and of eliminating agasketed joint.

• Special High-Pressure Closures (Fig. 3.1D) The channel barrel and the tube sheetare generally forged. The removable channel cover is seated in place by hydrostaticpressure, while a shear ring subjected to shearing stress absorbs the end force. Forpressures above 6205 kPa (900 lbf/in2) these designs are generally more economicalthan bolted constructions, which require larger flanges and bolting as pressure in-creases in order to contain the end force with bolts in tension. Relatively light-gaugeinternal pass partitions are provided to direct the flow of tube-side fluids but aredesigned 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 inFig. 3.1M; Types L and N are also used. All removable-bundle designs (except for the Utube) have a floating-head cover directing the flow of tube-side fluid at the floating tubesheet.


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 bydrawing the outer tube onto the inner one or by shrink fitting. These are usedwhere corrosive nature of the tube side fluid is such that no one metal or alloy iscompatible 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 inoutside diameter (1 inch= 25.4 mm). Wall thickness is measured in Birminghamwire gauge (BWG) units. The most commonly used tubes in chemical plants andpetroleum refineries are 19- and 25-mm (3/4- and 1-in) outside diameter. Standardtube 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 suchthat the tubing is produced to either average or minimum wall thickness. Seamlesscarbon steel tube of minimum wall thickness may vary from 0 to 20 percent above thenominal wall thickness. Average-wall seamless tubing has an allowable variation ofplus 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 ismade to minimum wall specifications.

Common practice is to specify exchanger surface in terms of total external squarefeet of tubing. The effective outside heat-transfer surface is based on the length oftubes measured between the inner faces of tube sheets. In most heat exchangersthere is little difference between the total and the effective surface. Significantdifferences 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 accordancewith 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 onboth sides of the tube. Fig. 5.7shows some of the commonly used fins. The fins aregenerally 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) Serratedradial fins. (c) Studded surface. (d) Joint between tubesheet and low fin tube with three timesbare surface. (e) External axial fins. ( f ) Internal axial fins. (9) Finned surface with internalspiral 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, isbeing used in shell-and-tube heat exchangers. The fins are radially extrudedfrom thick-walled tube to a height of 1.6 mm (1/16 in) spaced at 1.33 mm (19fins per inch) or to a height of 3.2 mm (1/8 in) spaced at 2.3 mm (11 fins perinch). External surface is approximately 2 1/2 times the outside surface of abare tube with the same outside diameter. Also available are 0.93-mm- (0.037-in-) high fins spaced 0.91 mm (28 fins per inch) with an external surface about3.5 times the surface of the bare tube. Bare ends of nominal tube diameter areprovided, while the fin height is slightly less than this diameter. The tube canbe inserted into a conventional tube bundle and rolled or welded to the tubesheet 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 inFig. 11-39. Internally finned tubes have been manufactured but have limitedapplication.

(b) Longitudinal fins are commonly used in double-pipe exchangers upon theoutside of the inner tube. U-tube and conventional removable tube bundlesare also made from such tubing. The ratio of external to internal surfacegenerally is about 10 or 15:1.

(c) Transverse fins upon tubes are used in low-pressure gas services. The primaryapplication 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 notpermit the use of a single alloy for the tubes, bimetallic (or duplex) tubes may beused. These can be made from almost any possible combination of metals. Tubesizes and gauges can be varied. For thin gauges the wall thickness is generallydivided equally between the two components. In heavier gauges the more expensivecomponent may comprise from a fifth to a third of the total thickness.

The component materials comply with applicable ASTM specifications, but aftermanufacture the outer component may increase in hardness beyond specificationlimits, and special care is required during the tube-rolling operation. When theharder material is on the outside, precautions must be exercised to expand thetube properly. When the inner material is considerably softer, rolling may not bepractical unless ferrules of the soft material are used.

In order to eliminate galvanic action the outer tube material may be stripped fromthe tube ends and replaced with ferrules of the inner tube material. When the endof a tube with a ferrule is expanded or welded to a tube sheet, the tube-side fluidcan contact only the inner tube material, while the outer material is exposed to theshell-side fluid. Bimetallic tubes are available from a small number of tube millsand 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 andridges. These are applicable in condenser and.

(b) The surface is prepared with special coating to provide a large number ofnucleation sites for use in boiling operations.

3.4.4 Tube arrangement

The tubes in an exchanger are usually arranged in an equilateral triangular, aquare orrotated square pattern see fig.3.14.

The triangular and rotated square pattern give higher heat transfer rates, but at theexpenses of higher pressure drop than the the square pattern. Square or rotated squareare used for hihger fouling fluid, where it is necessary to mechanically clean the outsideof the tubes. The recommend tube pitch is Pt = 1.25do. Where square pattern is usedfor easer of cleaning, the recommended minimum clearance between the tubes is 0.25 in(6.4 mm)


3.5 Shell side construction 47

p t

do

Square pitch

p t

Equilateral triangular pitch

pt

d o

Rotaed square

Flow

Figure 3.14. Tube patterns.

3.4.5 Tube side passes

The fluid in the tube is usually directed to flow back and forth in a number of passesthrough groups of tube arranged in parallel to increase the length of the flow path. Thenumber of passes is selected to give the required side design velocity. Exchangers are builtform one to up to 16 passes. The tube are arranged into the number of passes required bydividing up the exchanger headers (channels) with partition plates (pass partition) Thearrangement of the pass partition for 2,4 and 6 are shown in fig.3.19

1

2

1

2

3

4

1

2 3

45

6

Two tube passes

Four tube passes

Six tube passes

1

2 3

45

6

Figure 3.15. Tube arrangement: showing pass-partitions in headers.

3.5 Shell side construction

3.5.1 Shell Sizes

Heat-exchanger shells are generally made from standard- wall steel pipe in sizes up to305-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. Clearancesbetween the outer tube limit and the shell are discussed elsewhere in connection with thedifferent types of construction.



3.5.2 Shell-Side Arrangements

1. The one-pass shell (Fig. 3.1E) is the most commonly used arrangement. Con-densers from single component vapors often have the nozzles moved to the centerof the shell for vacuum and steam services. Solid longitudinal baffle is provided toform 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 twoshells in series.

3. For split flow (Fig. 3.1G), the longitudinal baffle may be solid or perforated. Thelatter feature is used with condensing vapors.

4. double-split-flow design is shown in Fig. 3.1H. The longitudinal baffles may besolid or perforated.

5. The divided flow design (Fig. 3.1J), mechanically is like the one-pass shell ex-cept for the addition of a nozzle. Divided flow is used to meet low-pressure-droprequirements. The kettle reboiler is shown in Fig. 3.1K. When nucleate boiling isto be done on the shell-side, this common design provides adequate dome space forseparation of vapor and liquid above the tube bundle and surge capacity beyondthe 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 generallyconstitute the most expensive component of the exchanger and are the one most likely tocorrode. Tube sheets, baffles, or support plates, tie rods, and usually spacers completethe bundle.

3.6.2 Baffle

Baffles are used to direct the side and tube side flows so that the fluid velocity is increasedto obtain higher heat transfer rate and reduce fouling deposits. In horizontal units baffleare used to provide support against sagging and vibration damage. There are differenttypes 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. Thedouble segmental baffle reduces crossflow velocity for a given baffle spacing. Thetriple segmental baffle reduces both cross-flow and long-flow velocities and has beenidentified as the window-cut baffle.


3.6 Baffles and tube bundles 49

a

b

c

d

Figure 3.16. Types of baffle used in shell and tube heat exchanger. (a) Segmental. (b)Segmental and strip. (c) Disc and doughnut. (d) Oriffice.

Minimum baffle spacing is generally one-fifth of the shell diameter and not lessthan 50.8 mm (2 in). Maximum baffle spacing is limited by the requirement toprovide adequate support for the tubes. The maximum unsupported tube spanin inches equals 74d0.75 (where d is the outside tube diameter in inches). Theunsupported tube span is reduced by about 12 percent for aluminum, copper, andtheir alloys.

Baffles are provided for heat-transfer purposes. When shell-side baffles are notrequired 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 sothat every pair of baffles will support each tube. Tube bundles are generally pro-vided with baffles cut so that at least one row of tubes passes through all the bafflesor support plates. These tubes hold the entire bundle together. In pipe-shell ex-changers with a horizontal baffle cut and a horizontal pass rib for directing tube



side flow in the channel, the maximum baffle cut, which permits a minimum of onerow of tubes to pass through all baffles, is approximately 33 percent in small shellsand 40 percent in larger pipe shells.

Maximum shell-side heat-transfer rates in forced convection are apparently obtainedby cross-flow of the fluid at right angles to the tubes. In order to maximize thistype of flow some heat exchangers are built with segmental-cut baffles and with notubes in the window (or the baffle cutout). Maximum baffle spacing may thus equalmaximum unsupported-tube span, while conventional baffle spacing is limited toone-half of this span.

The maximum baffle spacing for no tubes in the window of single segmental bafflesis unlimited when intermediate supports are provided. These are cut on both sidesof the baffle and therefore do not affect the flow of the shell-side fluid. Each supportengages all the tubes; the supports are spaced to provide adequate support for thetubes.

2. Rod Baffles Rod or bar baffles (fig. 3.17) have either rods or bars extendingthrough the lanes between rows of tubes. A baffle set can consist of a baffle withrods in all the vertical lanes and another baffle with rods in all the horizontal lanesbetween the tubes. The shell-side flow is uniform and parallel to the tubes. Stagnantareas do not exist.

One device uses four baffles in a baffle set. Only half of either the vertical or thehorizontal tube lanes in a baffle have rods. The new design apparently provides amaximum shell-side heat-transfer coefficient for a given pressure drop.

Figure 3.17. Rod baffles.

3. Impingement Baffle The tube bundle is customarily protected against impinge-ment by the incoming fluid at the shell inlet nozzle when the shell-side fluid is at ahigh velocity, is condensing, or is a twophase fluid. Minimum entrance area aboutthe nozzle is generally equal to the inlet nozzle area. Exit nozzles also require ade-quate area between the tubes and the nozzles. A full bundle without any provisionfor shell inlet nozzle area can increase the velocity of the inlet fluid by as much as300 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 toprotect 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 capabove the shell.

Impingement baffles or flow-distribution devices are recommended for axial tube-side nozzles when entrance velocity is high.

(a)(B)

(c) (d)

Figure 3.18. Impingment baffless;(a)Flat plate (b)curved plate (c)expanded or flared nozzle(d) jacket type.

4. Longitudinal Flow Baffles In fixed-tube-sheet construction with multipass shells,the baffle is usually welded to the shell and positive assurance against bypassingresults. Removable tube bundles have a sealing device between the shell and thelongitudinal baffle. Flexible light-gauge sealing strips and various packing deviceshave been used. Removable U-tube bundles with four tube-side passes and twoshell-side passes can be installed in shells with the longitudinal baffle welded inplace.

In split-flow shells the longitudinal baffle may be installed without a positive sealat the edges if design conditions are not seriously affected by a limited amount ofbypassing.

Fouling in petroleum-refinery service has necessitated rough treatment of tube bun-dles during cleaning operations. Many refineries avoid the use of longitudinal baffles,since the sealing devices are subject to damage during cleaning and maintenanceoperations.

3.6.3 Vapor Distribution

Relatively large shell inlet nozzles, which may be used in condensers under low pressureor 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 aminimum. The most significant bypass stream is generally between the outer tube limitand the inside of the shell. The clearance between tubes and shell is at a minimum forfixed-tube-sheet construction and is greatest for straight-tube removable bundles. Ar-rangements to reduce tube-bundle bypassing include:



1. Dummy tubes. These tubes do not pass through the tube sheets and can belocated close to the inside of the shell.

2. Tie rods with spacers. These hold the baffles in place but can be located toprevent bypassing.

3. Sealing strips. These longitudinal strips either extend from baffle to baffle or maybe inserted in slots cut into the baffles.

4. Dummy tubes or tie rods with spacers may be located within the pass partitionlanes (and between the baffle cuts) in order to ensure maximum bundle penetrationby the shell-side fluid.

When tubes are omitted from the tube layout to provide entrance area about animpingement plate, the need for sealing strips or other devices to cause properbundle 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 orpipe placed on the rods to locate the baffles. Occasionally baffles are welded to the tierods, and spacers are eliminated. Properly located tie rods and spacers serve both to holdthe bundle together and to reduce bypassing of the tubes.

In very large fixed-tube-sheet units, in which concentricity of shells decreases, baffles areoccasionally 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 corrosionor in vibratory service, in which metal baffles may cut the tubes.

Tube plate

baffle

Spacer

Rods

Figure 3.19. Baffle spacers and tie rods.

3.6.6 Tubesheets

Tubesheets are usually made from a round flat piece of metal with holes drilled for thetube ends in a precise location and pattern relative to one another. Tube sheet materialsrange as tube materials. Tubes are attached to the tube sheet by pneumatic or hydraulicpressure or by roller expansion. Tube holes can be drilled and reamed and can be machinedwith one or more grooves. This greatly increases the strength of the tube joint.



0.4mm

3 mm

ab c

Figure 3.20. Tube sheet joint

The tubesheet is in contact with both fluids and so must have corrosion resistance al-lowances and have metalurgical and electrochemical properties appropriate for the fluidsand velocities. Low carbon steel tube sheets can include a layer of a higher alloy metalbonded to the surface to provide more effective corrosion resistance without the expenseof using the solid alloy. The tube hole pattern or pitch varies the distance from one tubeto the other and angle of the tubes relative to each other and to the direction of flow. Thisallows the manipulation of fluid velocities and pressure drop, and provides the maximumamount of turbulance and tube surface contact for effective heat transfer. Where thetube and tube sheet materials are joinable, weldable metals, the tube joint can be furtherstrengthened by applying a seal weld or strength weld to the joint. A strength weld hasa 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 theshell and tube liquids from intermixing. In this treatment, the tube is flush with the tubesheet surface. The weld does not add metal, but rather fuses the two materials. In caseswhere it is critical to avoid fluid intermixing, a double tube sheet can be provided. In thisdesign, the outer tube sheet is outside the shell circuit, virtually eliminating the chanceof fluid intermixing. The inner tube sheet is vented to atmosphere so any fluid leak iseasily detected.

Mechanisms of attaching tubes to tube sheet

• Rolled Tube Joints Expanded tube-to-tube-sheet joints are standard. Properlyrolled joints have uniform tightness to minimize tube fractures, stress corrosion,tube-sheet ligament pushover and enlargement, and dishing of the tube sheet. Tubesare expanded into the tube sheet for a length of two tube diameters, or 50 mm (2in), or tube-sheet thickness minus 3 mm (1/8 in). Generally tubes are rolled for thelast of these alternatives. The expanded portion should never extend beyond theshell-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). Theflare or bell-mouth tube end is usually restricted to water service in condensers andserves 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 177C (350F), tube-sheet holes without grooves arestandard. For all other services with expanded tubes at least two grooves in eachtube hole are common. The number of grooves is sometimes changed to one or threein proportion to tube-sheet thickness.

• Expanding the tube into the grooved tube holes provides a stronger joint butresults in greater difficulties during tube removal (see Fig. 3.20a).



• Welded Tube Joints When suitable materials of construction are used, the tubeends may be welded to the tube sheets. Welded joints may be seal-welded for addi-tional tightness beyond that of tube rolling or may be strength-welded. Strength-welded joints have been found satisfactory in very severe services. Welded jointsmay 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 inthe tube hole. Fouling during normal operation followed by maintenance operationswill leave various impurities in and near the tube ends. Satisfactory welds are rarelypossible under such conditions, since tube-end welding requires extreme cleanlinessin the area to be welded.

• Tube expansion after welding has been found useful for low and moderate pres-sures. In high-pressure service tube rolling has not been able to prevent leakageafter weld failure.

• Double-Tube-Sheet Joints This design prevents the passage of either fluid intothe other because of leakage at the tube-to-tubesheet joints, which are generally theweakest points in heat exchangers. Any leakage at these joints admits the fluid tothe gap between the tube sheets. Mechanical design, fabrication, and maintenanceof double- tube-sheet designs require special consideration (see Fig. 3.20c).


55

4 Basic Design Equations of Heat ExchangersThere are two types of design problems: sizing and rating. In sizing the main objectiveis to find the geometry of the heat exchanger. Rating is to find the duty or performancefor a given geometry.

RATING SIZINGGiven: 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. Theseare

• LMTD-method,

• NTU-ε-method and

• θ-method.

These notation are explained in the respective sections.

4.1 LMTD-MethodAssumptions

• Steady state flow (mh,mc)

• Constant overall heat transfer coefficient (U)

• Constant specific heat (Cph, Cpc)

• negligible heat loss to surrounding

Heat Transfer (or rate equation)

Q = UA∆TlmF (4.1)

where

Q = heat transferred per unit time W (duty)U = overall heat transfer coefficientA = heat transfer area∆Tlm = logarithmic mean temperature differenceF = temperature correction factor


56 4 Basic Design Equations of Heat Exchangers

4.1.1 Logarithmic mean temperature different

∆Tlm =∆T2 −∆T1

ln(∆T2/∆T1)(4.2)

The temperature difference ∆T1, ∆T2 for different tube heat exchanger are defined below:

ThoThi

Tci Tco

Thi

Tho

Tco

Tci

∆T1 ∆T2

Cocurrent

ThoThi

Tco Tci

Thi

ThoTco

Tci

∆T1

∆T2

Counter current

Thi

Tho

Tco

Tci

∆T1

Thi

Tho

Tci

Tco

Tc

Shell and Tube

Figure 4.1. Temperature distribution

∆T1 ∆T2

Cocurrent Thi − Tci Tho − Tco

Counter current Thi − Tco Tho − Tci

Shell and tube Thi − Tco Tho − Tci

Plate heat exchanger Thi − Tco Tho − Tci

Example 1 water at a rate of 68 kg/min is heated from 35 to 65 oC by an oil having aspecific heat of 1.9 kJ/kg oC. The oil enters the exchanger at 110 oC and leaves at 75 oC.Calculate the logarithmic mean temperature difference for

1. counter current

2. co-current

Solution


4.1 LMTD-Method 57

ThoThi

Tci Tco

Thi=110 Co

Tho=75 Co

Tco=65 Co

Tci= 35 Co

∆T =751 ∆T =10 C2o

Cocurrent

ThoThi

Tco Tci

∆T1=45 Co

∆T C2=40o

Counter current

Thi=110 Co

Tci= 35 CoTci= 35 Co

Tco=65 Co

Tho=75 Co

Figure 4.2. Temperature distribution

1. counter current (see Fig.4.2)

∆Tlm =∆T2 −∆T1

ln(∆T2/∆T1)=

10− 75

ln(10/75)= 32.26oC (4.3)

2. co-current (see Fig.4.2)

∆Tlm =∆T2 −∆T1

ln(∆T2/∆T1)=

40− 45

ln(40/45)= 42.45oC (4.4)

4.1.2 Correction Factor

• For double pipe heat exchangerF = 1 (4.5)

• Shell and tube heat exchanger. For a 1 shell 2 tube pass exchanger the correctionfactor is given by:

F =

√(R2 + 1) ln

[1−S

1−RS

]

(R− 1) ln

2−S

[R+1−

√(R2+1)

]

2−S

[R+1−

√(R2+1)

]

(4.6)

where

R =T1 − T2

t2 − t1, S =

t2 − t1T1 − t1

(4.7)

or in words

R =Range of shell f luid

Range of tube fluid, S =

Range of tube fluid

Maximum temperature difference(4.8)

the derivation of the equation 4.6 is given by Kern (1950). The equation can beused for any exchanger with an even number of tube passes and is plotted in Fig.4.4.The correction factor for 2 shell passes and 4 or multiple of 4 tube passes is

F =

[R2+1

2(R−1)

]ln 1−S

1−RS

ln2/S−1−R+(2/S)

√(1−S)(1−RS)+

√R2+1

2/S−1−R+(2/S)√

(1−S)(1−RS)−√R2+1

(4.9)

These equations are plotted on fig.4.4



Example 1 For example calculate the correction factor for

1. 1-2 shell and tube heat exchanger and

2. 2-4 shell and tube heat exchanger

using the equation and the graph.

T1 = 35oC, T2 = 65oC, t1 = 110oC, t2 = 75oC

R =T1 − T2

t2 − t1=

35− 65

75− 110= 0.86, S =

t2 − t1T1 − t1

=75− 110

35− 110= 0.467 (4.10)

From the graph of fig.4.4

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

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

T2

t1

t2

1-2 Shell and TubeT1 t1

T1

T2

t2

2-4 Shell and Tube

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


4.1 LMTD-Method 59

Figure 4.4. Temperature correction factor: one shell, 2 shell pass, divide flow shell and splitflow shell and cross flow

4.1.3 Overall heat transfer coefficient

Typical values of the overall heat transfer coefficient for various types of heat exchnagerare given in . More expensive data can be found in in

The determination of U is often tedious and needs data not yet available in preliminarystages of the design. Therefore, typical values of U are useful for quickly estimating therequired surface area. The literature has many tabulations of such typical coefficients forcommercial 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 valuesare for unfavorable conditions such as lower flow velocities, higher viscosities, and addi-tional fouling resistances. Higher values are for more favorable conditions. Coefficientsof actual equipment may be smaller or larger than the values listed. Note that the val-ues should not be used as a replacement of rigorous methods for the final design of heatexchangers, although they may serve as a useful check on the results obtained by these



methods.

Table 4.1. Typical overall coefficient

Hot Fluid Cold fluid U (W/m2 oC)

Heat exchangersWater Water 800-1500Organic solvents organic solvent 100-300light oils light oils 100-400heavy oils heavy oils 50-300Gases gass 10-50CoolersOrganic solvents water 250-750light oils water 350-900heavy oils water60-900gase water 20-300organic solvent brine 150-500water brine 600-1200Gases Brine 15-250

HeatersSteam Water 1500-4000Steam organic solvent 500-1000Steam light oils 300-900Steam heavy oils 60-450Steam gass 30-300Dowtherm Heavy oils 50-300Dowtherm Gases 20-200flue gases steam 30-100flue gases hydrocarbon vapor 30-100

CondensersAqueous vapor water 1000-1500Organic vapor Water 700-1000Organic (some non condensable gases) Water 500-700Vacuum condensers Water 200-500

VaporizersSteam Aqueuos solutions 1000-1500Steam Light organics 900-1200Steam Heavy organics 600-900

Alternatively the overall heat transfer coefficient is evalauted from the individual heattransfer coefficient as:

1

Uo

=1

ho

+1

hod

+do ln (do/di)

2kw

+do

di

1

hi

+do

di

1

hid

(4.11)


4.2 ε- NTU 61

where

Uo = the overall coefficient based on the outside area of the tubeW/m2 oCho = outside fluid film coefficient,W/m2 oChi = inside fluid film coefficient,W/m2 oChod = outside dirt coefficient (Fouling factor),W/m2 oChi = inside dirt coefficient,W/m2 oCkw = thermal conductivity of the tube wall material, W/moCdo = tube outside diameter,mdi = tube inside diameter,m

4.1.4 Heat transfer coefficient

The heat transfer coefficient is governed by general function for forced convective as

Nu =hd

k= f

(Re, Pr,

d

L,

µ

µw

)(4.12)

and for natural convection as

Nu =hd

k= f

(Gr, Pr,

µ

µw

)(4.13)

Design equations for the heat transfer coefficient for various flow geometry (tube, plate)and configuration are given in Appendix 1. Design equation for the heat transfer coefficientfor condensation and boiling is given also in appendix A.

4.1.5 Fouling factor (hid, hod)

Heat transfer may be degraded in time by corrosion, deposits of reaction products, or-ganic growths, etc. These effects are accounted for quantitatively by fouling resistances.Extensive data on fouling factor are given TEMA standards. Typical fouling factors forcommon process and service fluids are given in the table 4.2. These values are for shelland tube heat exchangers with plain (not finned) tubes.

4.2 ε- NTUThe effectiveness (ε) of a heat exchanger is defined as the ratio between the actual heatload to the maximum possible heat load.

ε =Q

Qmax

(4.14)

This is related to the heat exchanger size and capacity as

ε = f(NTU,C) (4.15)

Where NTU is number of transfer unit and is defined as

NTU = N =UA

Cmin

(4.16)

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

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



Table 4.2. Fouling factorFluid Coefficient (W/m2 oC) Factor (resistance (m2 oC/W )

River water 3000-12000 0.003-0.0001Sea water 1000-3000 0.001-0.0003cooling water (towers) 3000-6000 0.0003-0.00017Towns water (soft) 3000-5000 0.0003-0.0002Towns water (hard) 1000-2000 0.001-0.0005Steam condensate 1500-5000 0.00067-0.0002Steam oil free 4000-10000 0.0025-0.00001Steam oil traces 2000-5000 0.0005-0.0002Refrigerated brine 3000-5000 0.0003-0.0002Air and industrial gases 5000-10000 0.0002-0.00001Flue gases 2000-5000 0.0005-0.0002Organic vapor 5000 0.0002Organic liquids 5000 0.0002Light hydrocarbons 5000 0.0002Heavy hydrocarbons 2000 0.0005Boiling organics 2500 0.0004Condensing organics 5000 0.0002Heavy transfer fluids 5000 0.0002Aqueous salt solutions 3000-5000 0.0003-0.0002

MhCph < McCpc ⇒ Cmin = MhCph, Cmax = McCpc (4.18)

MhCpc > McCpc ⇒ Cmin = McCpc, Cmax = MhCph (4.19)

Qmax = Cmin(Thi − Tci) (4.20)

C =Cmin

Cmax

(4.21)

εh =Thi − Tho

Thi − Tci

, εc =Tco − Tci

Thi − Tci

(4.22)

ε =∆Tc

Tspan

(4.23)

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


4.2 ε- NTU 63

Tci

Tco

Thi

ThoTspan

∆θ

0 A

Figure 4.5. Temperature distribution in counter current flow

The ε equation for various heat exchanger configuration is given as

• Parallel flow

ε =1− exp [−N(1 + C)]

1 + C(4.24)

• Counter current flow

ε =1− exp [−N(1 + C)]

1− C exp [−N(1− C)](4.25)

• Cross flow

1. Both fluid unmixed mixed

ε = 1− exp

[exp(−NCn)− 1

Cn

](4.26)

wheren = N−0.22 (4.27)

2. Both fluid mixed

ε =

[1

1− exp(−N)− 1+

C

1− exp(−NC)− 1− 1

N

]−1

(4.28)

3. Cmax mixed, Cmin unmixed

ε =1

C1− exp [−C (1− exp(−N))] (4.29)



4. Cmax unmixed, Cmin mixed

ε = 1− exp− 1

C[1− exp(−NC)]

(4.30)

• One shell pass, 2,4, 6 tube passes

ε = 2

1 + C +√

(1 + C2)1 + exp

[−N

√(1 + C2)

]

1− exp[−N

√(1 + C2)

]

−1

(4.31)

• Condenserε = 1− e−N (4.32)

• Evaporatorε = 1− e−N (4.33)

Alternatively these equations are presented in a graphical form. The various curves of εvs NTU can be found in textbooks like Kern (1964( and Perry and Green (2000).

4.3 Link between LMTD and NTU• Cocurrent

ln(

∆T1

∆T2

)= ln

(Thi − Tci

Tho − Tco

)= Nh + Nc (4.34)

• Counter current

ln(

∆T1

∆T2

)= ln

(Thi − Tco

Tho − Tci

)= Nh −Nc (4.35)

4.4 The Theta MethodAlternative method of representing the performance of heat exchangers may be given byTheta method [146] as

Θ =∆Tm

Tspan

(4.36)

where ∆Tm is the mean temperature difference and Tspan is the maximum temperaturedifference (Thi−Tci) (see Fig. 4.5). The Theta method is related is related to the associatedε and NTU methods by expressions

Θ =∆Tm

Tspan

=ε

NTU(4.37)

The relationship between parameters are often presented in graphical form as shown inFig.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 anymultiple 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 agiven heat exchanger. Remembering that the primary duty is to perform its thermal dutywith the lowest cost yet provide excellent in service reliability, the selection of fluid streamallocations should be of primary concern to the designer. There are many trade-offs influid allocation in heat transfer coefficients, available pressure drop, fouling tendenciesand operating pressure.

• The higher pressure fluid normally flows through the tube side. With their smalldiameter and nominal wall thicknesses, they are easily able to accept high pressuresand avoids more expensive, larger diameter components to be designed for highpressure. If it is necessary to put the higher pressure stream in the shell, it shouldbe placed in a smaller diameter and longer shell.

• Place corrosive fluids in the tubes, other items being equal. Corrosion is resistedby using special alloys and it is much less expensive than using special alloy shellmaterials. Other tube side materials can be clad with corrosion resistant materialsor epoxy coated.

• Flow the higher fouling fluids through the tubes. Tubes are easier to clean usingcommon mechanical methods.

• Because of the wide variety of designs and configurations available for the shellcircuits, such as tube pitch, baffle use and spacing, multiple nozzles, it is best toplace 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 providingincreased available surface.

Quiz: The top product of a distillation column is condensed using sea water. Allocatethe 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 andcause 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 insertsare sometimes used to reduce erosion at the tube inlet. Typical design velocity are givenbelow:

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; thelower values in the range given below will apply to molecular weight materials

Vacuum 50 to 70 m/sAtmospheric pressure 10 to 30 m/sHigh pressure 5 to 10 m/s

5.1.3 Stream temperature

The closer the temperature approach used (the difference between the outlet temperatureof one stream and the inlet temperature of the other stream) the larger will be the heattransfer area required for a given duty. The optimum value will depend on the applicationand can only be determined by making an economic analysis of alternative designs. Asa general guide the greater temperature difference should be at least 20 oC. and theleast temperature difference 5 to 7 oC for cooler using cooling water and 3 to 5 oC usingrefrigerated brine. The maximum temperature rise in recirculated cooling water is limitedto around 30 oC. Care should be taken to ensure that cooling media temperatures are keptwell above the freezing point of the process materials. When heat exchange is betweenprocess fluids for heat recovery the optimum approach temperatures will normally not belower than 20 oC.

5.1.4 Pressure drop

The value suggested below can be used as a general guide and will normally give designsthat are near the optimum.

Liquids

Viscosity<1 mN s/m2 ∆p< 35kN/m2

Viscosity=1 to 10mN s/m2 ∆p= 50-70 kN/m2

Gas and Vapors

High vacuum 0.4-0.8 kN/m2

Medium vacuum 0.1×absolute pressure1 to 2 bar 0.5×system gauge pressureAbove 10 bar 0.1×system gauge pressure

When a high-pressure drop is utilized, care must be taken to ensure that the resultinghigh fluid velocity does not cause erosion or flow -induced tube vibration.

5.1.5 Fluid physical properties

In the correlation used to predict heat-transfer coefficients, the physical properties areusually evaluated at the mean stream temperature. This is satisfactory when the tem-perature change is small, but can cause a significant error when change in temperatureis large. In these circumstances , a simple and safe procedure is to evaluate the heattransfer coefficients at the stream inlet and outlet temperatures and use the lowest of the


68 5 Thermal Design

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

Q =A [U2(T1 − t2)− U2(T2 − t1)]

ln[

U2(T1−t2)U1(T2−t1)

] (5.1)

where U1, U2 are evaluated at the end of the exchanger.

If the variation is too large for these simple methods to be used it will be necessaryto divide the temperature-enthalpy profile into sections and evaluate the heat transfercoefficients and area required for each section.

5.2 Design dataBefore discussing actual thermal design, let us look at the data that must be furnishedby 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 gasdensity is not furnished; it is not really necessary for liquids, as their properties donot vary with pressure.

4. allowable pressure drop for both streams. This is a very important parameter forheat exchanger design. Generally, for liquids, a value of 0.5-0.7 kg/cm2 is permittedper shell. A higher pressure drop is usually warranted for viscous liquids, especiallyin the tubeside. For gases, the allowed value is generally 0.05-0.2 kg/cm2, with 0.1kg/cm2 being typical.

5. fouling resistance for both streams. If this is not furnished, the designer shouldadopt values specified in the TEMA standards or based on past experience.

6. physical properties of both streams. These include viscosity, thermal conductivity,density, and specific heat, preferably at both inlet and outlet temperatures. Viscos-ity data must be supplied at inlet and outlet temperatures, especially for liquids,since the variation with temperature may be considerable and is irregular (neitherlinear nor log-log).

7. heat duty. The duty specified should be consistent for both the shellside and thetubeside.

8. type of heat exchanger. If not furnished, the designer can choose this based uponthe characteristics of the various types of construction described earlier. In fact, thedesigner 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 expandersor reducers. However, sizing criteria for nozzles are usually more stringent than forlines, especially for the shellside inlet. Consequently, nozzle sizes must sometimes beone size (or even more in exceptional circumstances) larger than the correspondingline sizes, especially for small lines.

10. preferred tube size. Tube size is designated as O.D., thickness, length. Some plantowners have a preferred O.D., thickness (usually based upon inventory considera-tions), and the available plot area will determine the maximum tube length. Manyplant owners prefer to standardize all three dimensions, again based upon inventoryconsiderations.


5.3 Tubeside design 69

11. maximum shell diameter. This is based upon tube-bundle removal requirementsand is limited by crane capacities. Such limitations apply only to exchangers withremovable tube bundles, namely U-tube and floating-head. For fixed-tubesheetexchangers, the only limitation is the manufa’s fabrication capability and the avail-ability of components such as dished ends and flanges. Thus, floating-head heatexchangers are often limited to a shell I.D. of 1.4-1.5 m and a tube length of 6 mor 9 m, whereas fixedtubesheet heat exchangers can have shells as large as 3 m andtubes lengths up to 12 m or more.

12. materials of construction. If the tubes and shell are made of identical materials, allcomponents should be of this material. Thus, only the shell and tube materials ofconstruction need to be specified. However, if the shell and tubes are of differentmetallurgy, the materials of all principal components should be specified to avoidany 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 operatingscenarios, and whether operation is continuous or intermittent.

5.3 Tubeside designTubeside calculations are quite straightforward, since tubeside flow represents a simplecase of flow through a circular conduit. Heat-transfer coefficient and pressure drop bothvary with tubeside velocity, the latter more strongly so. A good design will make the bestuse 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 leadto a certain velocity. Usually, this velocity is unacceptably low and therefore has to beincreased. By incorporating pass partition plates (with appropriate gasketing) in thechannels, the tubeside fluid is made to flow several times through a fraction of the totalnumber of tubes. Thus, in a heat exchanger with 200 tubes and two passes, the fluid flowsthrough 100 tubes at a time, and the velocity will be twice what it would be if there wereonly 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 Prandtlnumber, and the tube diameter. These can be broken down into the following fundamen-tal parameters: physical properties (namely viscosity, thermal conductivity, and specificheat); tube diameter; and, very importantly, mass velocity.

The variation in liquid viscosity is quite considerable; so, this physical property has themost dramatic effect on heat-transfer coefficient. The fundamental equation for turbulentheat-transfer inside tubes is:

Nu = CReaPrb

(µ

µw

)c

, (5.2)

or

h = Ck

D

(GD

µ

)a (Cpµ

k

)b(

µ

µw

)c

(5.3)


70 5 Thermal Design

whereNu = hde

kNusselt number

Pr = Cpµk

Prandtl number

Re ρudµ

Reynolds number

de4AP

hydraulic diameterA cross-sectional areaP wetted perimeteru fluid velocityµw fluid viscosity at the tube wall temperaturek fluid thermal conductivityCp fluid specific heat

C =

0.021 gases0.023 non-viscous liquid0.027 viscous liquid

a = 0.8b = 0.3 for coolingb = 0.4 for heatingc = 0.14

Viscosity influences the heat-transfer coefficient in two opposing ways- as a parameter ofthe Reynolds number, and as a parameter of Prandtl number. Thus, from Eq. 5.3:

h ∝ µ0.8−0.33 = µ0.47 (5.4)

In other words, the heat-transfer coefficient is inversely proportional to viscosity to the0.47 power. Similarly, the heat-transfer coefficient is directly proportional to thermalconductivity to the 0.67 power.

These two facts lead to some interesting generalities about heat transfer. A high thermalconductivity promotes a high heat-transfer coefficient. Thus, cooling water (thermalconductivity of around 0.55 kcal/hmC) has an extremely high heat-transfer coefficientof typically 6,000 kcal/hm2C, followed by hydrocarbon liquids (thermal conductivitybetween 0.08 and 0.12 kcal/hmC) at 250-1,300 kcal/hm2C, and then hydrocarbon gases(thermal conductivity between 0.02 and 0.03 kcal/hmC) at 50-500 kcal/hm2C.

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 towardthe upper limit of the range for hydrocarbon liquids.

The range of heat-transfer coefficients for hydrocarbon liquids is rather large due to thelarge variation in their viscosity, from less than 0.1 cP for ethylene and propylene to morethan 1,000 cP or more for bitumen. The large variation in the heat-transfer coefficientsof hydrocarbon gases is attributable to the large variation in operating pressure. Asoperating pressure rises, gas density increases. Pressure drop is directly proportional tothe square of mass velocity and inversely proportional to density. Therefore, for the samepressure drop, a higher mass velocity can be maintained when the density is higher. Thislarger mass velocity translates into a higher heat-transfer coefficient.

5.3.2 Pressure drop

The pressure drop due to friction exists because of the shear stress between the fluid andthe 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 givenas

−(

dp

dz

)

f

=4τo

d=

4fG2

2dρ, (5.5)

where G is the mass flux in kg/m2s and f is the friction factor calculated using a Blasius-type model as

f =

0.3164Re0.25 Re ≥ 2320

64Re

Re < 2320 .

Integration of equation B.1 yields

∆p =4fG2

2ρ

L

d, (5.6)

Mass velocity strongly influences the heat-transfer coefficient. For turbulent flow, thetubeside heat-transfer coefficient varies to the 0.8 power of tubeside mass velocity, whereastubeside pressure drop varies to the square of mass velocity. Thus, with increasing massvelocity, 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 toincrease mass velocity further.

Furthermore, very high velocities lead to erosion. However, the pressure drop limitationusually becomes controlling long before erosive velocities are attained. The minimumrecommended 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 agiven tubeside flow rate, the pressure drop rises to the cube of this increase. In actualpractice, the rise is somewhat less because of lower friction factors at higher Reynoldsnumbers, so the exponent should be approximately 2.8 instead of 3.

Tubeside pressure drop rises steeply with an increase in the number of tube passes. Con-sequently, it often happens that for a given number of tubes and two passes, the pressuredrop is much lower than the allowable value, but with four passes it exceeds the allowablepressure drop. If in such circumstances a standard tube has to be employed, the designermay be forced to accept a rather low velocity. However, if the tube diameter and lengthmay be varied, the allowable pressure drop can be better utilized and a higher tubesidevelocity realized.

The following tube diameters are usually used in the CPI: (1/4, 3/8, 1/2, 5/8, 3/4, 1, 11/4, 1 1/2 in. Of these, 3/4 in. and 1 in. are the most popular. Tubes smaller than 3/4in. O.D. should not be used for fouling services. The use of small-diameter tubes, such as1 in., is warranted only for small heat exchangers with heat-transfer areas less than 20-30m2.

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 aparticular 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, apressure drop of 0.7 kg/cm2 per shell is permitted for liquid streams. If there are fivesuch preheat exchangers, a total pressure drop of 3.5 kg/cm2 for the circuit would bepermitted. If the pressure drop through two of these exchangers turns out to be only 0.8kg/cm2, the balance of 2.7 kg/cm2 would be available for the other three.


72 5 Thermal Design

5.4 Shell side designShell side design The shellside calculations are far more complex than those for the tube-side. This is mainly because on the shellside there is not just one flow stream but oneprincipal cross-flow stream and four leakage or bypass streams. There are various shell-side flow arrangements, as well as various tube layout patterns and baffling designs, whichtogether determine the shellside stream analysis.

5.4.1 Shell configuration

TEMA defines various shell patterns based on the flow of the shellside fluid through theshell: 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 leavesfrom the other end. This is the most common shell type - more heat exchangers are builtto 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 throughone-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 shortof 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 leavesat 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 alarge temperature cross can be achieved.

A TEMA J shell is a divided-flow shell wherein the shellside fluid enters the shell at thecenter and divides into two halves, one flowing to the left and the other to the right andleaving separately. They are then combined into a single stream. This is identified as aJ 1-2 shell. Alternatively, the stream may be split into two halves that enter the shell atthe two ends, flow toward the center, and leave as a single stream, which is identified asa J 2-1 shell.

A TEMA G shell is a split-flow shell (see Figure 3.1). This construction is usually em-ployed for horizontal thermosyphon reboilers. There is only a central support plate andno baffles. A G shell cannot be used for heat exchangers with tube lengths greater than3 m, since this would exceed the limit on maximum unsupported tube length specified byTEMA - 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 shellis basically two G shells placed side-by-side, so that there are two full support plates. Thisis 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. Theadvantage of G and H shells is that the pressure drop is drastically less and there are nocross baffles.

A TEMA X shell (see Figure 3.1) is a pure cross-flow shell where the shellside fluid entersat the top (or bottom) of the shell, flows across the tubes, and exits from the opposite sideof the shell. The flow may be introduced through multiple nozzles located strategicallyalong the length of the shell in order to achieve a better distribution. The pressure dropwill be extremely low - in fact, there is hardly any pressure drop in the shell, and whatpressure drop there is, is virtually all in the nozzles. Thus, this configuration is employedfor cooling or condensing vapors at low pressure, particularly vacuum. Full support platescan be located if needed for structural integrity; they do not interfere with the shellsideflow 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), rotatedtriangular (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 turbulenceand therefore a high heat-transfer coefficient. However, at the typical tube pitch of 1.25times the tube O.D., it does not permit mechanical cleaning of tubes, since access lanesare not available. Consequently, a triangular layout is limited to clean shellside services.For services that require mechanical cleaning on the shellside, square patterns must beused. Chemical cleaning does not require access lanes, so a triangular layout may be usedfor dirty shellside services provided chemical cleaning is suitable and effective.

A rotated triangular pattern seldom offers any advantages over a triangular pattern, andits use is consequently not very popular.

For dirty shellside services, a square layout is typically employed. However, since this is anin-line pattern, it produces lower turbulence. Thus, when the shellside Reynolds numberis low (< 2,000), it is usually advantageous to employ a rotated square pattern becausethis produces much higher turbulence, which results in a higher efficiency of conversionof pressure drop to heat transfer.

As noted earlier, fixed-tubesheet construction is usually employed for clean services onthe shellside, Utube construction for clean services on the tubeside, and floating-headconstruction for dirty services on both the shellside and tubeside. (For clean serviceson both shellside and tubeside, either fixed-tubesheet or U-tube construction may beused, although U-tube is preferable since it permits differential expansion between theshell and the tubes.) Hence, a triangular tube pattern may be used for fixed-tubesheetexchangers and a square (or rotated square) pattern for floating-head exchangers. ForU-tube exchangers, a triangular pattern may be used provided the shellside stream isclean and a square (or rotated square) pattern if it is dirty.

5.4.3 Tube pitch

Tube pitch is defined as the shortest distance between two adjacent tubes.


74 5 Thermal Design

For a triangular pattern, TEMA specifies a minimum tube pitch of 1.25 times the tubeO.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 patternsis either 1.25 times the tube O.D. or the tube O.D. plus 6 mm, whichever is larger. Forexample, 20-mm tubes should be laid on a 26-mm (20 mm + 6 mm) square pitch, but25-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 tothe smallest shell diameter for a given number of tubes. However, in exceptional cir-cumstances, the tube pitch may be increased to a higher value, for example, to reduceshellside 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 bemaintained for the shellside fluid, and prevent failure of tubes due to flow-induced vibra-tion. There are two types of baffles: plate and rod. Plate baffles may be single-segmental,double-segmental, or triple-segmental, as shown in Figure 5.2.

Figure 5.2. Types of baffles.

Baffle spacing. Baffle spacing is the centerline-to-centerline distance between adjacentbaffles. It is the most vital parameter in STHE design.

The TEMA standards specify the minimum baffle spacing as one-fifth of the shell insidediameter or 2 in., whichever is greater. Closer spacing will result in poor bundle pene-tration by the shellside fluid and difficulty in mechanically cleaning the outsides of thetubes. Furthermore, a low baffle spacing results in a poor stream distribution as will beexplained later.

The maximum baffle spacing is the shell inside diameter. Higher baffle spacing willlead to predominantly longitudinal flow, which is less efficient than cross-flow, and largeunsupported tube spans, which will make the exchanger prone to tube failure due toflow-induced vibration.

Optimum baffle spacing. For turbulent flow on the shellside (Re > 1,000), the heat-transfer coefficient varies to the 0.6-0.7 power of velocity; however, pressure drop varies


5.4 Shell side design 75

to the 1.7-2.0 power. For laminar flow (Re < 100), the exponents are 0.33 for the heat-transfer coefficient and 1.0 for pressure drop. Thus, as baffle spacing is reduced, pressuredrop 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 diameterthat 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 ineach baffle to permit the shellside fluid to flow across the baffle. This is expressed as apercentage of the shell inside diameter. Although this, too, is an important parameterfor 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 theshellside due to large deviation from an ideal situation, as illustrated in Figure 5.4.

Figure 5.4. Effect of small and large baffle cuts.

It is strongly recommended that only baffle cuts between 20% and 35% be employed. Re-ducing baffle cut below 20% to increase the shellside heat-transfer coefficient or increasingthe baffle cut beyond 35% to decrease the shellside pressure drop usually lead to poor de-signs. Other aspects of tube bundle geometry should be changed instead to achieve thosegoals. For example, doublesegmental baffles or a divided-flow shell, or even a cross-flowshell, may be used to reduce the shellside pressure drop.


76 5 Thermal Design

For single-phase fluids on the shellside, a horizontal baffle cut (Figure 5.5) is recommended,because this minimizes accumulation of deposits at the bottom of the shell and alsoprevents stratification. However, in the case of a two-pass shell (TEMA F), a vertical cutis 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 (thatis, 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 - preferablywithin 20%

of each other. If they differ by more than that, repeated acceleration and deceleration takeplace along the length of the tube bundle, resulting in inefficient conversion of pressuredrop 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 fourleakage or bypass streams, as illustrated in Figure 5.6. Tinker (4) proposed calling thesestreams the main cross-flow stream (B), a tube-to-baffle-hole leakage stream (A), a bundlebypass 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 otherstreams are not as effective. The A stream is fairly efficient, because the shellside fluidis in contact with the tubes. Similarly, the C stream is in contact with the peripheraltubes around the bundle, and the F stream is in contact with the tubes along the pass-partition lanes. Consequently, these streams also experience heat transfer, although ata lower efficiency than the B stream. However, since the E stream flows along the shellwall, 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 aparticular set of exchanger geometry and shellside flow conditions by any sophisticatedheatexchanger thermal design software. Essentially, the five streams are in parallel andflow along paths of varying hydraulic resistances. Thus, the flow fractions will be such thatthe pressure drop of each stream is identical, since all the streams begin and end at theinlet 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 isestablished.


78 5 Thermal Design

Besides these methods there is some proprietary methods putout by various organizationfor use by their member companies. A number of these method are based on one of theabove methods. Some are based upon a judicious combination of methods 3 and 4 aboveand supplemented by further research data. Among the most popular of the proprietarymethods, judged by their large clientele are

• Heat Transfer Research Inc. (HTRI), Alliambra, california. This method is alsoknown 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 inCoulson and Richardson’s

5.4.8 Heat transfer coefficient

Nu = 0.36Re0.55Pr1/3

(µ

µw

)0.14

, (5.7)

whereNu = hde

kNusselt number

Pr = Cpµk

Prandtl numberRe = Gde

µReynolds number

de = 4AP

hydraulic diameterA = cross-sectional flow areaP = wetted perimeterG = M

AsMass flux

As = (pt−do)DslBpt

fluid viscosity at the tube wall temperature

pt = pitch diameterDs = shell diameterlB = Baffle spacing

Hydraulic diameter (Fig. 5.1)

de =

p2t−πd2

o/4

πdofor square pitch

0.87p2t /2−πd2

o/8

πdo/2for equilateral triangular pitch

5.4.9 Pressure drop

∆p = 4f(

Ds

d

) (ρu2

2

) (L

lb

) (µ

µw

)−0.14

, (5.8)

where

f =

0.3164Re0.25 Re ≥ 2320

64Re

Re < 2320 .

L=tube length

lB = baffle spacing. The term (L/lB) is the number of times the flow crosses the tubebundle=(NB + 1). Where NB is the number of baffles.


5.5 Design Algorithm 79

5.5 Design Algorithm

Step1SpecificationDefine duty Q

Make energy balance if neededto calcualted unspecified flow

rates or temperatureQ=M )c pc c2 c1 h ph h1 h2c (T -T )=M C (T -T

Step2Calculate physical properties

Step3Assume value of overall

coefficient Uo,ass

Step 4Decide number of shell and

tube passesCalculate T∆ ∆T , F andlm m

Step 5Determine heat transfer area

required A To o,ass=q/U ∆ m

Step 6Decide type, tube size, material,

layoutAssign fluids to shell or tube

Step 7Calculate number of tubes

Step 8Calculate shell diameter

Step 9Estimate tube-side heat

transfer coefficient

Step 10Decide baffle spacing and estimateshell side heat transfer coefficient

Step 11Calculate overall heat transfer

Coefficient including fouling factorsUo,cal

Step 12Estimate tube and shell side

pressure drop

Step 13Estimate cost of heat exchanger

Can design beoptmized toreduce cost?

Accept design

Is pressure dropswithin specification?

0<(Uo,cal o,ass o,ass-U )/U <30Set Uo,ass=Uo,cal

Yes

No

No

Figure 5.7. Design procedure for shell and tube heat exchanger.


80 6 Specification sheet

6 Specification sheetSpecification sheet is a data sheet that contains the information provided by the customerto the vendor for the process and mechanical designs of an exchanger. After the processdesign is done, the engineer fills in some further information. The rest of the informationis filled after the mechanical design is completed. The specification sheet is a medium ofcommunication between different parties involved in the procurement, design and fabri-cation of heat exchanger. It is also used to compare the performance of the installed unitwith the design conditions.

6.1 Information includedThe information contained in the sheet is best decribed by a data sheet. Although eachcompany has its own version of data sheet, the most popular one is that of the TEMAstandards. 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 includedThe 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 inthe specification sheet.

6.3 Operation conditionsThe following operating conditions regarding the exchanger operation should be knownto 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 thecomposition of the streams, their thermal and physical properties and any phasechange 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 datafrom each bid should be put on a large data sheet. During evaluation the following factorshould be kept in mid:

1. The design submitted by the bidders should meet the heat transfer and pressuredrop 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 bechecked 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 ortheir 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 balanceand be unduly favorable to the bidder


82 6 Specification sheet

Figure 6.1. Data sheet


83

7 Storage, Installation, Operation and MaintenanceProper storage, installation handling and correct start up emergency, and shutdown pro-cedure are important for the successful working of a well designed and fabricated heatexchanger. regular cleaning, maintenance and repairs are necessary to ensure trouble freeoperation of the unit for its designed life span. These will be discussed in the followingsections.

NOTE: Before placing your equipment in operation, environment and service conditionsshould be checked for compatibility with materials of construction. Contact your nearestheat exchanger Standard representative if you are not sure what the actual materials ofconstruction are.

Successful performance of heat transfer equipment, length of service and freedom fromoperating 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 repairsand 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 StorageStandard heat exchangers are protected against the elements during shipment. If theycannot be installed and put into operation immediately upon receipt at the jobsite, cer-tain precautions are necessary to prevent deterioration during storage. Responsibility forintegrity of the heat exchangers must be assumed by the user. The manufacturer will notbe responsible for damage, corrosion or other deterioration of heat exchanger equipmentduring 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. Thefollowing suggested practices are provided solely as a convenience to the user, who shallmake his own decision on whether to use all or any of them.


84 7 Storage, Installation, Operation and Maintenance

1. On receipt of the heat exchanger, inspect for shipping damage to all protective cov-ers. If damage is evident, inspect for possible contamination and replace protectivecovers 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 toprevent rusting or contamination.

3. Heat exchangers for oil service, made of ferrous materials, may be pressure-testedwith oil at the factory. However, the residual oil coating on the inside surfaces ofthe exchanger does not preclude the possibility of rust formation. Upon receipt,fill these exchangers with appropriate oil or coat them with a corrosion preventioncompound for storage. These heat exchangers have a large warning decal, indicatingthat they should be protected with oil.

4. The choice of preservation of interior surfaces during storage for other service appli-cations depends upon your system requirements and economics. Only when includedin the original purchase order specifications will specific preservation be incorporatedprior to shipment from the factory.

5. Remove any accumulations of dirt, water, ice or snow and wipe dry before movingexchangers into indoor storage. If unit was not filled with oil or other preservative,open drain plugs to remove any accumulated moisture, then reseal. Accumulationof moisture usually indicates rusting has already started and remedial action shouldbe taken.

6. Store under cover in a heated area, if possible. The ideal storage environment forheat exchangers and accessories is indoors, above grade, in a dry, low humidity at-mosphere which is sealed to prevent entry of blowing dust, rain or snow. Maintaintemperatures between 70F and 105F (wide temperature swings may cause con-densation and ”sweating” of steel parts). Cover windows to prevent temperaturevariations caused by sunlight. Provide thermometers and humidity indicators atseveral 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 assilica gel), or portable dehumidifiers, to remove moisture from the air in the storageenclosure. Thermostatically controlled portable heaters (vented to outdoors) maybe required to maintain even air temperatures inside the enclosure.

8. Inspect heat exchangers and accessories frequently while they are in storage. Starta log to record results of inspections and maintenance performed while units arein storage. A typical log entry should include, for each component, at least thefollowing:

(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 toseal off the space between the floating tubesheet and inside shell diameter to applythe test in paragraph

11. Consult your nearest sales representative for reference drawings showing installationof 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 rollingdepth should be tubesheet thickness minus 1/8”.

• Do not re-roll tubes that are not leaking since this needlessly thins the tubewall.

13. It is recommended that when a heat exchanger is dismantled, new gaskets be usedin reassembly.

• Composition gaskets become brittle and dried out in service and do not providean effective seal when reused.

• Metal or metal jacketed gaskets in initial compression match the contact sur-faces and tend to work-harden and cannot be recompressed on reuse.

14. Use of new bolting in conformance with dimension and ASTM specifications of theoriginal design is recommended where frequent dismantling is encountered. CAU-TION: Do not remove channel covers, shell covers, floating head covers or bonnetsuntil all pressure in the heat exchanger has been relieved and both shell side andtube side are completely drained.

15. If paint deterioration begins, as evidenced by discoloration or light rusting, considertouch-up or repainting. If the unit is painted with our standard shop enamel, areasof light rust may be wire brushed and touched-up with any good quality air-dryingsynthetic enamel. Units painted with special paints (when specified on customers’orders) may require special techniques for touch-up or repair. Obtain specific infor-mation from the paint manufacturer. Painted steel units should never be permittedto rust or deteriorate to a point where their strength will be impaired. But a lightsurface rusting, on steel units which will be re-painted after installation, will notgenerally cause any harm. (See Items 3 and 4 for internal surface preservation.)

16. If the internal preservation (Items 3 and 4 ) appears inadequate during storage,consider additional corrosion prevention measures and more frequent inspections.Interiors coated with rust preventive should be restored to good condition and re-coated promptly if signs of rust occur.

7.2 Installation

7.2.1 Installation Planning

1. On removable bundle heat exchangers, provide sufficient clearance at the stationaryend to permit the removal of the tube bundle from the shell. On the floating headend, 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 permitremoval and replacement of tubes and at the other end provide sufficient clearanceto permit tube rolling.



3. Provide valves and bypasses in the piping system so that both the shell side andtube 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 fromthe 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 toprevent 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 settleand cause piping strains. Foundation bolts should be set accurately. In concretefootings, pipe sleeves at least one pipe size larger than the bolt diameter slipped overthe bolt and cast in place are best for this purpose as they allow the bolt centers tobe adjusted after the foundation has set.

8. Install proper liquid level controls and relief valves and liquid level and temperaturealarms, etc.

9. Install gauge glasses or liquid level alarms in all vapor or gas spaces to indicate anyfailure occurring in the condensate drain system and to prevent flooding of the heatexchanger.

10. Install a surge drum upstream from the heat exchanger to guard against pulsationof fluids caused by pumps, compressors or other equipment.

11. Do not pipe drain connections to a common closed manifold; it makes it moredifficult 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 toinstallation. Make sure it is thoroughly cleaned to remove all preservation materialsunless stored full of the same oil being used in the system, or the coating is solublein the lubricating system oil. If the exchanger was oil-tested by any Standard andyour purchase order did not specify otherwise, the oil used was Tectyl 754, a light-bodied oil which is soluble in most lubricating oils. Where special preservations wereapplied, you should consult the preservative manufacturer’s product informationdata for removal instructions.

2. If the heat exchanger is not being stored, inspect for shipping damage to all pro-tective covers upon receipt at the jobsite. If damage is evident, inspect for possiblecontamination 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 canbe 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 toinstalling. Do not expose internal passages of the heat exchanger to the atmospheresince moisture or harmful contaminants may enter the unit and cause severe damageto the system due to freezing and/or corrosion.

5. After piping is complete, if support cradles or feet are fixed to the heat exchanger,loosen foundation bolts at one end of the exchanger to allow free movement. Over-sized holes in support cradles or feet are provided for this purpose.


7.3 Operation 87

6. If heat exchanger shell is equipped with a bellows-type expansion joint, removeshipping supports per instructions.

7.3 Operation

1. Be sure entire system is clean before starting operation to prevent plugging of tubesor shell side passages with refuse. The use of strainers or settling tanks in pipelinesleading to the heat exchanger is recommended.

2. Open vent connections before starting up.

3. Start operating gradually. See Table 1 for suggested start-up and shut-down proce-dures for most applications. If in doubt, consult the nearest manufactuerer repre-sentative for specific instructions.

4. After the system is completely filled with the operating fluids and all air has beenvented, close all manual vent connections.

5. Re-tighten bolting on all gasketed or packed joints after the heat exchanger hasreached operating temperatures to prevent leaks and gasket failures. Standard pub-lished torque values do not apply to packed end joints.

6. Do not operate the heat exchanger under pressure and temperature conditions inexcess of those specified on the nameplate.

7. To guard against water hammer, drain condensate from steam heat exchangers andsimilar 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 vibrationand will result in reduced operating life.

10. Under no circumstances is the heat exchanger to be operated at a flowrate greaterthan that shown on the design specifications. Excessive flows can cause vibrationand severely damage the heat exchanger tube bundle.

11. Heat exchangers that are out of service for extended periods of time should beprotected against corrosion as described in the storage requirements for new heatexchangers. Heat exchangers that are out of service for short periods and use wateras 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 heatexchanger on a daily basis to prevent stagnant water conditions that can ultimatelyprecipitate corrosion.



1. Clean exchangers subject to fouling (scale, sludge deposits, etc.) periodically, de-pending on specific conditions. A light sludge or scale coating on either side of thetube greatly reduces its effectiveness. A marked increase in pressure drop and/orreduction in performance usually indicates cleaning is necessary. Since the difficultyof cleaning increases rapidly as the scale thickens or deposits increase, the intervalsbetween cleanings should not be excessive.

2. Neglecting to keep tubes clean may result in random tube plugging. Consequentoverheating or cooling of the plugged tubes, as compared to surrounding tubes, willcause physical damage and leaking tubes due to differential thermal expansion ofthe metals.


7.3 Operation 89

3. To clean or inspect the inside of the tubes, remove only the necessary tube sidechannel covers or bonnets, depending on type of exchanger construction.

4. If the heat exchanger is equipped with sacrificial anodes or plates, replace these asrequired.

5. To clean or inspect the outside of the tubes, it may be necessary to remove the tubebundle. (Fixed tubesheet exchanger bundles are non-removable).

6. When removing tube bundles from heat exchangers for inspection or cleaning, ex-ercise care to see that they are not damaged by improper handling.

• The weight of the tube bundle should not be supported on individual tubesbut should be carried by the tubesheets, support or baffle plates or on blockscontoured to the periphery of the tube bundles.

• Do not handle tube bundles with hooks or other tools which might damagetubes. Move tube bundles on cradles or skids.

• To withdraw tube bundles, pass rods through two or more of the tubes andtake the load on the floating tubesheet.

• Rods should be threaded at both ends, provided with nuts, and should passthrough a steel bearing plate at each end of the bundle.

• Insert a soft wood filler board between the bearing plate and tubesheet face toprevent 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 returnthrough another tube.

• A hardwood spreader block must be inserted between the cable and eachtubesheet to prevent damage to the tube ends.

7. If the heat exchanger has been in service for a considerable length of time withoutbeing removed, it may be necessary to use a jack on the floating tubesheet to breakthe bundle free.

• Use a good-sized steel bearing plate with a filler board between the tubesheetface and bearing plate to protect the tube ends.

8. Lift tube bundles horizontally by means of a cradle formed by bending a light-gaugeplate 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. Avoidany 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 listedbelow:

• Circulating hot wash oil or light distillate through tube side or shell side willusually 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 beeffective in removing more stubborn deposits. Use in accordance with themanufacturer’s instructions.

11. Some tubes have inserts or longitudinal fins and can be damaged by cleaning whenmechanical means are employed. Clean these types of tubes chemically or consultthe nearest manufacturer representative for the recommended method of cleaning.



• If the scale is hard and the above methods are not effective, use a mechanicalmeans. Neither the inside nor the outside of the tube should be hammeredwith a metallic tool. If it is necessary to use scrapers, they should not be sharpenough to cut the metal of the tubes. Take extra care when employing scrapersto prevent tube damage.

Do not attempt to clean tubes by blowing steam through individual tubes. Thisoverheats the individual tube and results in severe expansion strains and leakingtube-to-tubesheet joints.

12. Table 2 shows safe loads for steel rods and eyebolts.


91

8 Heat exchanger tube side mainenance (Repair vs

replacement(This subject of chapter is collected from: Bruce W Schafer Framatome ANP, Inc. 155Mill Ridge Road Lynchburg, VA 24502 (434) 832-3360 [email protected])

Abstract The traditional method of repairing degraded tubes in shell-and-tube heatexchangers is to remove the effected tubes from service by plugging. Since heat exchangersare designed with excess heat transfer capability, approximately 10% of tubes can beplugged before performance is affected. When the number of plugged tubes becomesexcessive, heat exchanger efficiency is lost, resulting in reduced power output, high systempressure drop, further heat exchanger damage, or abnormal loads placed on other plantheat exchangers.

As an option to component retubing or replacement, repair methods, including tube sleev-ing and tube expansion, have proven to be an effective method to repair defective tubesand keep the existing heat exchanger in service. For the sleeving process, a new tubesection is installed inside the existing tube to bridge across the degraded area. Tubeexpansion is used to close off a gap between the tube and the tubesheet or end plate (toeliminate a leak path) or between the tube and tube support (to minimize vibration).While not all heat exchangers can be returned to their original design condition by per-forming tube repairs, in some instances it may be possible to get many more years ofuseful life out of a heat exchanger at a fraction the cost of replacement.

This paper presents options which the Plant Maintenance Engineer should consider inmaking the repair versus replacement decision. This includes the repair options (sleevingand tube expansion), other conditions within the heat exchanger, and the effect of tuberepair on heat exchanger performance.

8.1 IntroductionTraditionally, when maintenance is performed on shell-and-tube heat exchangers, theonly options considered when tube defects are found are to plug tubes and, when thenumber of plugs became too great, replace the heat exchanger. The decision to replacethe heat exchanger was based on a number of factors. These included: the number oftubes plugged, the number of forced outages due to tube damage (and the cost associatedwith replacing lost power and repairing the damaged tubes), the impact that the pluggedheat 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 heatexchanger, and the expected life of the unit (how much longer will the unit operate beforeretirement).

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 heatexchanger due to damage in one area is an expensive as well as a schedule and manpowerintensive option.

The typical means for major heat exchanger repair included complete replacement, re-bundling, and retubing, as described below.

• For the replacement option, the entire heat exchanger shell and tube bundle arereplaced with a new unit.

• For rebundling, the shell is temporarily removed from the heat exchanger and theold tube bundle, including, at a minimum, tubes, tube supports, and tubesheet, areremoved. A new tube bundle is inserted and the shell is welded back in place.

• For retubing, either the shell (u-tube design) or tube side access cover (straight


92 8 Heat exchanger tube side mainenance (Repair vs replacement

tubes) is removed from the heat exchanger and the old tubes are removed from thebundle. New tubes are then inserted and re-attached to the tubesheet (typically byeither mechanical expansion, welding, or both). In many instances, the existing shellside hardware is used as-is, although some modifications may be made. Retubingis typically performed on straight tube heat exchangers, such as condensers andcoolers.

Since the 1970’s, tube sleeving has been used to allow damaged tubes to remain in ser-vice. The sleeves are installed by various means (roll, explosive, or hydraulic expansion,explosively welded, or press-fit or epoxied in place) over the defective area of the tube.Through the use of sleeving, which is a low-cost option to retubing, rebundling, or re-placement, the useful life of a heat exchanger can be economically extended. The decisionto perform sleeving also can be made with short notice as opposed to replacement (2-6weeks compared with 18 months), possibly allowing repairs to be performed the sameoutage that the damage is noted. Tube expansion also can be performed to minimize oreliminate leakage within heat exchangers. In the tubesheet, tubes can be re-expanded tostrengthen the original tube-to-tubesheet joint, reducing or eliminating leakage and pro-longing the life of the heat exchanger. Expansions also can be made deep within the tubeto expand the tube into tube support plates and end plates. These expansion can reducetube-to-plate clearance for vibration control or, at end plates, to minimize steam flowfrom the high to low pressure side of the plate.Since the 1970’s, tube sleeving has beenused to allow damaged tubes to remain in service. The sleeves are installed by variousmeans (roll, explosive, or hydraulic expansion, explosively welded, or press-fit or epoxiedin 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 exchangercan be economically extended. The decision to perform sleeving also can be made withshort notice as opposed to replacement (2-6 weeks compared with 18 months), possiblyallowing repairs to be performed the same outage that the damage is noted.

Tube expansion also can be performed to minimize or eliminate leakage within heat ex-changers. In the tubesheet, tubes can be re-expanded to strengthen the original tube-to-tubesheet joint, reducing or eliminating leakage and prolonging the life of the heatexchanger. Expansions also can be made deep within the tube to expand the tube intotube support plates and end plates. These expansion can reduce tube-to-plate clearancefor vibration control or, at end plates, to minimize steam flow from the high to lowpressure side of the plate.

8.2 Repair vs. Replace - Factors To ConsiderThere are numerous factors to consider when deciding whether to repair the tubes in aheat exchanger or to perform a larger repair scope and rebundle or replace the component.The following factors should be considered when making the repair vs. replace decision.

1. The budget available for repair or replacement needs to be determined. Typically,the cost of performing a substantial heat exchanger repair (consisting of plug re-moval, tube inspection, tube expansion, and sleeving) is less than 10% of the cost ofreplacing the unit. Because of the lower cost, the payback time on the repair optionis 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 bepossible to justify a 3 repair to the unit in the near-term and a scheduled replacementwhen a longer outage can be planned. If there are a large number of tube plugsto remove, or if they are difficult to remove (explosive or welded), then the cost torepair the heat exchanger will increase, and the scheduled time needed on-site maynot 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 andsleeves installed, thereby minimizing the overall maintenance cost.

2. The location and quantity of the tube defects need to be examined to decide iftube repair is an option. Tube repair may be appropriate if the damage is limitedto 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 alonger sleeve (up to the full length of the tube) to ensure that all tube defects arerepaired. However, if the u-bend region of the tube is damaged then tube repair isnot possible. Also, it would not be possible to install a sleeve if a large portion ofthe tube had damage but there was inadequate clearance for a long sleeve at thetube end.

3. One of the more important items to consider when deciding whether a heat ex-changer can be repaired is the condition of the remainder of the heat exchanger.The condition of the shell side components, such as the impingement plates, tubesupports, end plates, and other structural members, should be in good shape if along term repair is being planned. An evaluation also should be made of the shellthickness in areas that are prone to shell erosion/corrosion. If the tube repair is onlya short-term fix, to allow component operation until a replacement heat exchangercan 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 torepair or replace a heat exchanger. If the only problem with the heat exchanger isin 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 itwill be very difficult to pay back the cost for replacement over the remaining plantlife.

5. The outage time required to repair a heat exchanger, even when tube and shell sideinspections 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. Thisallows other work to be performed in the vicinity of the heat exchanger. Alongwith 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 usuallybe performed through the manways and pass partition plates. Less repair equipmentis required, resulting in less space being needed in the area of the heat exchangerfor setup and storage. In addition, the time required to prepare for tube repair ismuch less than for replacement (2- 6 weeks compared with 18 months), allowing adecision on repair to be made just before, or even during, an outage.

6. At nuclear plants, the added cost for the disposal of radioactively contaminatedheat exchangers must be taken into account. Before disposal, there is the cost ofsurveying the heat exchangers for release and, if contamination is found, they musteither be decontaminated or disposed of as radioactive waste. Tube repairs caneliminate these costs.

7. If the heat exchanger is being replaced to eliminate detrimental materials in thecooling system (i.e. copper in the condensate/feedwater system) then tube sleevingwill not be beneficial. The only solution would be to retube/rebundle/replace tochange out the tube material.

8.3 Heat Exchanger maintenance OptionsThere have always been options available to either repair or replace heat exchanger tubesin 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-destructiveexaminations, such as eddy current testing, visual inspections, or leak tests) is to plug thetube. 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 insertedinto the tube to minimize future tube vibration damage.

At the beginning of the life of a heat exchanger, inserting a few plugs into damaged tubeshas little effect on the performance of the heat exchanger. However, if heat exchangerproblems continue, and the number of plugs increases significantly, it is possible thatthe heat exchanger will eventually reach a point that it will not handle the full loadthat is placed on it. This is due to a combination of loss of heat transfer area and theincreased pressure drop. In addition, as the number of plugged tubes increases, abnormaltemperature 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 demiseof the heat exchanger.

Once the number of plugs reaches a unacceptable level, the heat exchanger will need to berepaired, 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 transferarea. Also, the heat load from the bypassed heat exchanger will be transferred to anotherheat exchanger in the string, resulting in greater than normal operating flow rates andhigher degradation in that heater.

The following sections show the options that can be used to replace or repair the entireheat 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 toinstall sleeves over the defective portions of the tubes. The sleeve consists of a smallerdiameter piece of tubing that is inserted into the parent tube and positioned over thetube defects. After insertion, each end of the sleeve is expanded into the parent tubematerial. These expansions serve the dual function of structurally anchoring the sleeveinto the tube and providing a leak limiting path, allowing the sleeve to become the newpressure boundary for the tube. This means that a sleeved tube can have a 100% through-wall indication and still remain in-service, since the sleeve is now the new structural andpressure boundary. The installation of the sleeve into the tube will allow the majority ofthe 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 toprepare for future repair. It may be cost effective to plug a quantity of tubes, per the non-destructive examination results, each outage using a removable plug. When the quantityof 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 whileallowing 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 heatexchangers. After insertion, the full length of the sleeve is expanded into the parenttube. 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 fulllength expansion step, shown in Figure 8.2, the sleeve ends are trimmed flush with theexisting 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 whenthe tubes have multiple defects along their length. Full length sleeving is an attractiveoption 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 heatexchanger that is close to its as designed condition.

Full length sleeving is comparable in many ways to retubing in the methods employed toinstall the sleeves. However, since removal of the existing tube is not required, and thetypical number of tubes that will be full length sleeved are below the number that wouldbe retubed, the cost for material and manhours are much less than for retubing, makingsleeving a cost-effective option to return and keep tubes in service.

Figure 8.2. Full Length Sleeve Expansion

Partial Length Structural Sleeve

This type of sleeve is used to repair shorter defects in the tube. The sleeve can beinstalled anywhere along the straight length of the tube. Various methods are used toexpand the sleeve in place. These include roll expansion (both in the tubesheet and inthe 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 installationof 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 bemade. This expansion is similar to the original tube-to-tubesheet roll. Freespan rollexpansions are made to either a torque controlled setting or to a diameter controlledhardstop setting. Usually, freespan roll expansions are only used when the sleevelength is relatively short, since it can be difficult to insert a roll expander deep intothe tube. Both the tubesheet and freespan roll expansion parameters are set so thatthey can provide both the structural and leakage requirements for the sleeve.



• For sleeves installed deep within the tube, a hydraulic expansion device is used toconnect the sleeve to the tube. The expander consists of multiple plastic bladdersthat are filled with high pressure water. As the water pressure increases, the bladdersexpanded against the inside of the sleeve, pushing the sleeve into the tube. Theexpansion process, which is computer controlled, continues until either a presetvolume of water or a preset pressure is reached. At this point the sleeve is properlyexpanded and the bladders are depressurized. Hydraulic expansions can be madeanywhere along the tube length since the expander is connected to flexible highpressure tubing and is not restricted by tube end access. The expansion parametersare qualified to meet the proper structural and leakage requirements for the sleeve.

• Full length expansions are not usually used for structural or leak limiting purposesbut instead are used to improve heat transfer and flow through the sleeve and toclose the annulus between the sleeve and tube. The full length expansion is madeby placing a tool, with seals on each end, into the sleeve. The inside of the sleeveis filled and then pressurized with water to a preset pressure setting, expanding thesleeve 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 thestructural and leak limiting sleeve-to-tube joint.

Many times, the partial length structural sleeves are used to repair indications at oneparticular area of the tube, such as wear damage at tube support locations, crackingin 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 sectionof 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 toover 12 feet in length.

Qualification testing is performed on the structural sleeves to ensure that they can with-stand the design temperature and pressure conditions imposed on them. The test resultsmust show that the sleeve will be the new pressure boundary even with a 100% through-wall indication in the parent tube. Sleeves of this type, using mechanical expansions (rolland 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 barrierto tube end erosion. These sleeves are usually very thing, are not designed to act as apressure boundary or structural repair, and are installed in areas of high turbulence. Thematerials for these sleeves are compatible with the existing tube material and may includeplastic inserts. The sleeves are either roll or hydraulic expanded or pressed or epoxiedin place. If tube end erosion is occurring, or is expected to occur, the use of these tubeend sleeves will protect and prolong the life of the parent tube, although over time tubeerosion may begin to occur at the end of the sleeve. Many heat exchanger tube ends havebeen 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 needto be determined. If the defects are in one or a few short areas then either a single ora couple of partial length sleeves could be used. However, if the defects are spacedthroughout 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 supportdesignations must be clearly identified to insure that the sleeve is installed at thecorrect location along the tube length. This is especially true in areas where theremay 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 needsto be known. If there are u-bend defects that may require plugging then the tubeshould not be sleeved. Sleeving is an option if the remainder of the tube is in goodshape.

• The space available at the tube end to insert a sleeve and its installation toolingneeds to be known, as shown in Figure 8.4. If a short, partial length sleeve is beingused, the amount of space required is not as critical, although there can still beaccess issues around the tubesheet periphery for hemi-head channel covers and atpass partition plates. However, if a full length sleeve is required, there will need tobe a significant amount of clearance from the tubesheet face.

• Inspection records need to be reviewed to determine if there are any tube insidediameter (ID) restrictions that would block the sleeve from being inserted to thetarget location. The size of the eddy current probe used for the inspection, plus anyother hardware that has been inserted into the tube, can be used to help determinethe 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. Whilethe presence of the sleeve reduces the inside diameter of the tube, which will resultin the need for a smaller inspection probe, the probe will remain large enoughto detect pluggable tube indications (usually greater than 40%), however small



indications may go undetected. As part of the post-sleeve inspection, the sleeve andits attachment to the tube should be examined. There is no need to inspect thesection of the parent tube between the sleeve expansions since this is no longer partof the pressure boundary.

• If tube cleaning is to be performed in the heat exchanger, then the type of sleeve tobe installed needs to be evaluated. If on-line cleaning is performed, the sleeve sizecannot restrict the passage of the balls or brushes. For off-line cleaning, the projec-tiles need to pass through the sleeve without becoming stuck. Many sleeves that areinstalled in tubes that require cleaning are full length expanded to ensure the bestresults for the cleaning equipment. If it appears that tube sleeving is possible, theninformation 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 plugremoval.

• If it is expected that sleeving may be performed, then it is important that the propersleeve material be purchased in advance of the job.

• The sleeve material needs to be compatible with the heat exchanger parent tubingand with the water chemistry within the heat exchanger. The galvanic corrosionpotential between the sleeve and tube needs to be determined. Also, effects of crevicecorrosion between the sleeve and tube, in the heat exchanger water chemistry, needto be considered to determine if sleeving is a viable repair option.

• The sleeve dimensions need to fit the heat exchanger operating and design condi-tions plus any restrictions within the tube ID. The sleeve outside diameter (OD) isto be designed to fit into the tube but must be long enough to limit the amount ofsleeve expansion. The sleeve wall thickness needs to be sized for the heat exchangeroperating parameters, including any ASME Code minimum wall thickness calcula-tions, if needed. The sleeve length must be long enough to span the expected tubedefects but needs to be sized to fit any tube end clearance restrictions.

• Before installing sleeves into heat exchanger tubes, testing needs to be performed toset the installation parameters. Depending on the type of sleeve being used, thesetests may include setting the rolling torque, hydraulic expansion constants, and fulllength expansion pressure. In addition, depending on the application for the sleeve,there may be a need to do qualification testing, which would consist of hydrostaticleak and pressure tests and temperature and pressure cycling. These tests wouldverify 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 calculatethe heat transfer and flow loss due to sleeving. These calculations will give a sleeve-to-plug ratio that can be used to determine the expected improvement in heatexchanger performance after sleeving is complete (and tubes have been returned toservice, if applicable).

• The sleeve may need to be full-length expanded based on the heat exchanger oper-ating environment. However, the production rates for sleeve installation are lowerwhen full length expansions are performed. While full length expansion is typi-cally not needed in many applications, such as most feedwater heaters, it should beconsidered 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 heattransfer or flow



– if the tube-to-sleeve crevice needs to be eliminated in a hostile water chemistryenvironment

– if there are large eddy current probe fill factor restrictions

Figure 8.4. Required Clearance for Sleeve Installation

8.4.3 Tube Expansion

In addition to sleeving, it is possible to expand the tube to improve the heat exchangerperformance. These tube repairs can minimize further tube damage and maximize theuseful life of the heat exchanger. Two methods of tube expansion can be performed. Oneis to expand deep within the tube to close off a leak path between the tube and the endplate. The other is to re-expand the tube into the tubesheet to minimize tube-to-shellside leakage.

Tube-to-End Plate Expansion

In some heat exchangers, typically feedwater heaters, there are internal plates whichseparate one zone of the heat exchanger from another (usually condensing [steam] fromdrain cooler [liquid]). Due to the pressure differential across the plate, and the differenttemperatures and phases between the two sections, it is important that leakage not occurthrough 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 inFigure 8.5. When this occurs there is erosion of the end plate and tube vibration due tothe 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. Theleakage of steam also increases the drain cooler temperature, resulting in a less efficientheat exchanger. Expanding the tube can reduce the gap between the tube and the endplate. The expansion can be performed using either a roll or hydraulic expander. Once theexpander is in position the tube is expanded until it contacts the end plate. An accurateexpansion, which does not over-expand the tube into the plate (the tube needs to be ableto 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 hydraulicexpansion that automatically shuts off the pressurization system when it detects that thetube has contacted the plate.

After the tubes are expanded into the end plate, the steam flow is minimized or elimi-nated, reducing the drain cooler temperatures and increases plant efficiency. Further tubedamage, in the form of tube wear and adjacent tubes impacting on one another, will bereduced 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 withreplacement.

Tube-to-Tubesheet Expansion

In some heat exchanger designs, with a certain combination of materials, leaks developbetween the tube and tubesheet. In many low pressure units, the tube is only expandedinto the tubesheet, with no subsequent weld. Many of the leaks that occur in these unitsare the result of a fabrication error and can be corrected by re-expanding the joint tothe correct expansion size. However, leakage occasionally occurs in high pressure heatexchangers, typically feedwater heaters, even when the tubes have been welded to thetubesheet. The two prime causes of this leakage are in areas where the original tube-to-tubesheet weld has either cracked or eroded due to flow (in the case of soft materials, suchas 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 rollexpansion process. The expansion would increase the contact pressure between thetube and tubesheet, increasing the resistance to flow and decreasing or eliminating



leakage. This process could be performed on existing leaking tubes or preventativelyon all tubes in the tubesheet.

• If cracking is occurring at the original tube expansion transition it may be possibleto re-expand the tube deeper in the tubesheet (unless the cracking is occurringvery close to the shell side of the tubesheet). The tube would be expanded usinga qualified roll expansion process, to place the tube into tight contact with thetubesheet. This expansion would increase the contact pressure between the tube andtubesheet, increasing the resistance to flow and decreasing or eliminating leakage.This process could be performed either on existing leaking tubes or preventativelyon all tubes in the tubesheet.

Re-expanding tubes that either may be leaking or that could develop leaks in the futurecould significantly extend the life of an otherwise good heat exchanger. By re-expandingthe tubes, forced outages can be avoided and damage from the high pressure water spray-ing on adjacent tubes and on the shell will be eliminated. The cost to perform tubere-expansions will be minimal when compared with the cost of replacement heat exchang-ers and the cost of forced outages.

Items to Consider for Tube Expansion Repair

The following factors should be considered to determine if tube expansion is possible.

• The portion of the tube to be expanded needs to be determined.

– If leakage is occurring through the end plate, the expander will need to be longenough to reach the end plate location. The tube should be expanded using aprocess, such as hydraulic expansion, that will not lock the tube into the endplate. This expansion will not only reduce leakage through the plate but alsowill minimize future tube vibration due to the tight fit between the tube andplate.

– 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 fromthe tube to the shell side of the heat exchanger.

• The condition of the remainder of the tube needs to be known. If there are cracksalong the entire tube length then re-expanding the tube alone will not result in animprovement in heat exchanger performance.

• The space available at the tube end to insert the expansion tooling needs to beknown. 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 toolingis relatively short, although there can still be access issues around the tubesheetperiphery for hemi-head channel covers and at pass partition plates.

• For tube end plate expansions, the eddy current inspection records need to bereviewed to determine if there are any tube inside diameter restrictions that wouldblock the expander from being inserted to the end plate location. The size of theeddy current probe used for the inspection, plus any other hardware that has beeninserted 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 intothe tube, also need to be known so that tooling can be prepared to eliminate therestriction.


8.5 Replacement option 103

If it appears that tube expansion is possible, then information will be needed to ensurethat the heat exchanger is properly repaired. The following information is used whenplanning for tube expansion.

• Tube expansion will need to be coordinated with eddy current inspection and plugremoval.

• The tube expander design (diameter and length) needs to be based on the require-ments for the expansion. Before performing tube expansions into heat exchangertubes, testing needs to be performed to set the tooling operating parameters. De-pending on the type of expansion, these tests may include setting the rolling torquefor tubesheet re-expansions or setting the hydraulic expansion constants for endplate expansions. In addition, for the tube-intotubesheet re-expansion process, qual-ification testing should be performed. This would consist of hydrostatic leak andpressure tests and temperature and pressure cycling. These tests would verify thatthe 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 theheat exchanger are in good shape.

The old tubes are removed from the unit and new ones, typically manufactured froman improved material, are inserted, and then expanded, into place. Insertion of the newtubes is shown in Figure 8.6. In addition to performing retubing to replace damagedtubes, retubing has been performed to eliminate detrimental materials (such as copperfrom condenser tubes) to minimize damage to other equipment within the plant (nuclearsteam 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 unitsthe entire tube bundle, including tubes, tubesheet, and tube supports are replaced, asshown in Figure 8.7. The original shell and any other internal structural pieces wouldbe reused (although any necessary internal repairs could be made when the shell wasremoved). The new tube bundle can be manufactured to ensure that original designproblems with the existing unit are corrected. However, the same basic design mustbe 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 shownin 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 erosionor other problems may be occurring, to ensure that the current heat exchanger problemsdo not re-occur in the future. However, the cost associated with a full replacement is thegreatest of the three options, about 5% more than for rebundling . In addition, thereare no guarantees that the new heat exchanger design will not have new, unanticipatedproblems.

Figure 8.8. Heat Exchanger Replacement


8.6 Conclusions 105

8.6 ConclusionsThe costs associated with heat exchanger replacement can be significant. These costsinclude the new heat exchanger or tube bundle, the manpower required to remove theold and install the new heat exchanger components, plant modifications to allow for theremoval of the heat exchanger, and the amount of outage time associated with replace-ment. In addition, the replacement of a heat exchanger can adversely affect other workgoing on in the their vicinity. Because of the cost and time involved, and if the damageis 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 isconfined to one general area, there is a good possibility that the expense of a replacementcan be avoided. In addition, the time required to prepare for tube repair is much lessthan for replacement (2-6 weeks compared with 18 months), allowing a decision on repairto be made just before, or even into, an outage.

By removing plugs and installing sleeves, it is possible to return lost heat transfer area toservice. Tubes that would be likely to fail in the near term also can be repaired. This willimprove the performance and reliability of the heat exchanger. The cost to perform therepairs is also much less than for replacement (usually less than 1/10th the cost). Sleevinghas been shown to be a proven tube repair technique, having been performed since the1970’s. During this time, tube repairs have economically extended the useful life of heatexchangers worldwide.

As the number of plugged tubes approaches the upper limits or if damage is consistentlyoccurring in one area of a heat exchanger, tube repair, through both sleeving and tubeexpansions, should be considered to minimize future damage and extend the life of theheat

The following table shows the various heat exchanger repair options and the factors tobe considered when choosing each of the options. Note that the table contains selectedcriteria for evaluating component repair versus replacement options. A final decision toimplement a particular option should be made on a case by case basis with proper weightgiven to all factors. The information listed in this table is for relative comparison purposesonly.


106 9 Troubleshooting

9 Troubleshooting

9.1 Heat exchangers’ problemsHeat transfer equipment provides the economic and process viability for many plant op-erations. The basis for successful application of such equipment depends on the designer.The problem that should be anticipated by the design to avoid high maintenance orcleaning and costly shut down production include:

1. Fouling

2. Leakage

3. Corrosion

To anticipate maintenance problems the designer should need to be familiar with theplant location, process flow sheet, plant operation. Some of the questions that must beconsidered are:

1. will the heat exchanger need cleaning? how often? what cleaning method will beused?

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?willcycling 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 presentany 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 unresolvedproblem in heat transfer” ?

• the total cost of fouling - in highly industrialized nations - has been projected at0.25% 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 inthe process industries is now estimated at 6 billion ?

9.2.3 Types of Fouling

• Precipitation / Crystallization - dissolved inorganic salts with inverse solubility char-acteristics

• Particulate / Sedimentation - suspended solids, insoluble corrosion products, sand,silt

• Chemical Reaction - common in petroleum refining and polymer production

• Corrosion - material reacts with fluid to form corrosion products, which attach tothe 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 velocitygradients, oxygen depletion zones and crystal nucleation sites are established - afew minutes to a few weeks

• Migration - most widely studied phenomenon - involving tranport of foulant tosurface and various diffusion transport mechanisms

• Attachment - begins the formation of the deposit

• Transformation or Aging - another critical period when physical or chemical changescan increase deposit strenght and tenacity Removal or

• Re-entrainment - dependent upon deposit strength - removal of fouling layers bydissolution, 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 thannatural circulation calandria. This is to obtain a velocity of 10-15ft/sec. Al-though the cost of pumps and power added considerably to the cost of theequipment. This would be compared to the cost of production losses and costfor cleaning in order to arrive to at an economical design for a particular processapplication.

(b) Good shell side avoids eddies and dead zones where solid can accumulate. Inletand outlet connections should be located at the bottom and top of the shellside 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 isimportant, especially with cooling waters. Copper, copper alloy and stainlesssteels are satisfactory for most cooling waters

2. The fouling fluid should be inside tube. Hence easily removable flat cover plateswould be installed on the channel to facilitate cleaning if frequent physical cleaningis necessary. Horizontal installation would probably be chosen to avoid the cost ofscaffold 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 dueto sedimentation. For example spiral plate heat exchanger may be selected in placeof a multipass shell and tube heat exchanger unit to avoid settling of suspendedsolids in the shell side and at the bottom of the tube side bottom of the tube sidechannel.

9.2.7 Fouling cleaning methods

1. Chemical cleaning: Various chemicals (acids, chlorine) have been used to reducefouling and restore tube cleanliness. Acid may either be strong (which damage theequipment) or week (citric, formic, sulfamic) these are less effective. Acid cleaningis limited to once a year or less. The use of chlorine is being cutback or eliminatedin 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 (betweencleaning fouling builds up rapidly)

3. Rubber - ball cleaning: Automatic cleaning by means sponge -rubber balls is eco-nomical in areas where deposition, pollutants, chlorides and other corrodents exists.These ball distribute themselfs at random through the condenser, passing througha tube at an average of one every five minutes. slightly larger in diameter than thetube, 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 SurfaceLeaks may develop at

1. the tube-to-tubesheet joints of fixed tube sheet exchanger due

(a) to differential thermal expansion between the tube and shell causes overstress-ing of the rolled joints, or

(b) thermal cycling caused by frequent shutdowns or batch operation of the processmay cause the tubes to loosen in the tube holes.

2. Leaks may occur due to tube failure cause by vibration or differential thermal ex-pansion or dryout (for boilers and evaporators)

9.3.1 Cost of leakage

1. Large production losses or maintenance cost

2. Contamination of product:The leak/rupture of tubes leads to contamination or over-pressure of the low-pressure side. Failure to maintain separation between heat trans-fer and process fluids may lead to violent reaction in the heat transfer equipmentor in the downstream processing equipment.

9.3.2 Cause of differential thermal expansion

1. Unusual situation that lead to unexpected differential thermal expansion, for ex-ample, the tube side of a fixed-tube sheet condenser may be subjected to steamtemperature, with no coolant in the shell whenever a distillation column is steamedout in preparation for maintenance. Or an upset in the chemical process may subjectthe tubes to high temperatures

2. Start up at high temperature

3. Vibration (if the velocity at the inlet exceeded the critical velocity for two phaseflow)

4. Dryout of the tube cause by insufficient coolant or local overheating

Remedy of thermal expansion

1. Use of U tube or floating head instead of fixed tube sheet

2. Welding the tube to the tube sheet

3. Double tube sheet

4. Use large nozzle or vapor belts to give velocity well below the critical

To make the heat transfer process inherently safer, designers must look at possible in-teractions between heating/cooling fluids and process fluids. For relatively low-pressureequipment (<1000 psig), a complete failure of tubes may not be a credible overpressurescenario if the design pressure of the low-pressure side and associated equipment is greaterthan two-thirds of the design pressure of the high- pressure side (API RP 521 1993), or ifthe geometry of the tube layout is such that a complete break is not physically possible.

For high-pressure equipment (> 1000 psig), however, a complete failure should be consid-ered credible, regardless of pressure differential.



9.4 CorrosionThe heat transfer surface reacts chemically with elements of the fluid stream producinga 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, dissolvedoxygen, penetrating ions like chlorides and sulphates, the low pH and presence of variousother 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 candevelop stresses, which lead to stress corrosion.

• Overthinning: Expanding the tube into the tube sheet reduces the tube wall thick-ness and work-hardens the metal.

• The induced stresses can lead to stress corrosion.

Controlling Stress Corrosion Cracking

• Proper selection of the appropriate material.

• Remove the chemical species that promotes cracking.

• Change the manufacturing process or design to reduce the tensile stresses.

9.4.5 Galvanic corrosion

Galvanic corrosion is frequently referred to as dissimilar metal corrosion. Galvanic corro-sion can occur when two dissimilar materials are coupled in a corrosive electrolyte.


9.4 Corrosion 111

9.4.6 Pitting

Pitting is a localized form of corrosive attack. Pitting corrosion is typified by the formationof 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 theenvironment (pH, O2) Use more pitting resistant materials Improve the design of thesystem

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 atnarrow openings or spaces between two metal surfaces or between metals and nonmetalsurfaces.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 isthus created between the tube and the tube hole at this tube-sheet face. The tubeis 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, rearhead, nozzles,...) may be manufactured from same metal or dissimilar metals. Individualcomponents may be fabricated from single metal or bimetallic.

For the selection of material of construction, the corrosion chart must be consulted (Ap-pendix C of Coulson and Richardson [29]). The chart gives metal (alloy) vs chemical atvarious temperatures. Note:Before using the corrosion chart the notation given shouldread thoroughly.

9.4.10 Fabrication

Expanding the tube into the tube sheet reduces the tube wall thickness and work-hardensthe metal. The induced stresses can lead to stress corrosion. Differential expansionbetween tubes and shell in fixed-tube-sheet exchangers can develop stresses, which leadto stress corrosion.

When austenitic stainless-steel tubes are used for corrosion resistance, a close fit betweenthe tube and the tube hole is recommended in order to minimize work hardening and theresulting loss of corrosion resistance. In order to facilitate removal and replacement oftubes it is customary to roller-expand the tubes to within 3 mm of the shellside face ofthe tube sheet. A 3-mm- long gap is thus created between the tube and the tube holeat this tube-sheet face. In some services this gap has been found to be a focal point forcorrosion.

It is standard practice to provide a chamfer at the inside edges of tube holes in tube sheetsto prevent cutting of the tubes and to remove burrs produced by drilling or reaming thetube sheet. In the lower tube sheet of vertical units this chamfer serves as a pocketto collect material, dirt, etc., and to serve as a corrosion center. Adequate venting ofexchangers is required both for proper operation and to reduce corrosion.

Improper venting of the water side of exchangers can cause alternate wetting and dryingand accompanying chloride concentration, which is particularly destructive to the series300 stainless steels.

Certain corrosive conditions require that special consideration be given to complete drainagewhen the unit is taken out of service.

Particular consideration is required for the upper surfaces of tube sheets in vertical heatexchangers, for sagging tubes, and for shell-side baffles in horizontal units.

9.5 TroubleshootingThis chapter presents potential failure mechanisms for heat transfer equipment and sug-gests design alternatives for reducing the risks associated with such failures. The typesof 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 heattransfer equipment. Unless specifically noted, the failure scenarios apply to more thanone class of heat transfer equipment.


9.6 Past failure incidents 113

9.6 Past failure incidentsThis section provides several case histories of incidents involving failure of heat transferequipment to reinforce the need for the safe design practices presented in this chapter.

9.6.1 Ethylene Oxide Redistillation Column Explosion:

In March 1991, an Ethylene Oxide (EO) redistillation column exploded at a Seadrift,Texas chemical facility. The explosion was caused by energetic decomposition of essen-tially pure EO vapor and liquid mist inside the column.

A set of extraordinary circumstances was found to have coincided, resulting in the catalyticinitiation of decomposition in a localized region of a reboiler tube. Extensive investigationby reference [158] showed that:

1. A low liquid level in the column, plus a coinciding temporary condensate backupand accumulation of inert gas in the reboiler shell, significantly diminished the EOliquid fraction leaving the reboiler. Nevertheless, sufficient heat transfer capacity re-mained to satisfy the vaporization rate required by the column controls, so operationappeared 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 withtraces of water, high boiling glycols accumulated in the stalled tubes, increasingthe boiling point while reducing the heat flux and resulting mass flow rate. Thisself-reinforcing process continued leading to minimal EO vapor velocity through thestalled tubes. Since the vapor was no longer in equilibrium with boiling EO it couldmomentarily attain the 150oC temperature of the reboiler steam supply.

3. The insides of the reboiler tubes had collected a thin film of EO polymer containingpercent-level amounts of catalytic iron oxides. This film had in numerous placespeeled away from the tube wall producing a catalytic surface of low heat capacityand negligible effect on mass flow rate. EO vapor heating was aided by the absence ofliquid plus the small vapor velocity through the stalled tubes. These conditions ledto a rapid rate of film heating which encouraged a fast disproportionation reaction ofEO to predominate over slower polymerization reactions. The previously unknownfast reaction between EO vapor and supported high surface area iron oxide led to ahotspot and initiation of vapor decomposition. Once ignited the EO decompositionflame spread rapidly through the column causing overpressurization.

9.6.2 Brittle Fracture of a Heat Exchanger

An olefin plant was being restarted after repair work had been completed. A leak devel-oped on the inlet flange of one of the heat exchangers in the acetylene conversion preheatsystem. To eliminate the leak, the control valve supplying feed to the conversion systemwas shut off and the acetylene conversion preheat system was depressured. Despite thefact 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 resultedin extremely low temperature upon depressurization to atmospheric pressure.

The heat exchanger that developed the leak was equipped with bypass and block valvesto isolate the exchanger. After the leaking heat exchanger was bypassed, the acetyleneconversion system was repressured and placed back in service. Shortly thereafter, the firstexchanger 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 intwo fatalities, seven serious burn cases, and major damage to the olefins unit.

The acetylene converter pre-heater failed as a result of inadequate lowtemperature resis-tance during the low temperature excursion caused by depressuring the acetylene converter



system. The heat exchanger that failed was fabricated from ASTM A515 grade 70 car-bon steel. After the accident, all process equipment in the plant which could potentiallyoperate at less than 200F was reviewed for suitable low-temperature toughness [116].

Ed. Note: It should have been recognized that upstream cryogenic conditions may havea 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 oflow temperature plant streams and the plant refrigeration systems. This collection ofheat 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 columnrequired periodic heating and backflushing with methane to prevent excessive pressuredrop 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 anexplosion occurred which destroyed the cold box and disabled the ethylene plant for about5 months. An estimated 20 tons of hydrocarbon escaped. Fortunately, the hydrocarbondid not ignite.

An investigation revealed that the explosion was caused by the accumulation and sub-sequent violent decomposition of unstable organic compounds that formed at the lowtemperatures inside the cold box. The unstable ”gums55 were found to contain nitroand nitroso components on short hydrocarbon chains. The source of the nitrogen wasidentified as nitrogen oxides (NOx) present in a feed stream from a catalytic crackingunit. Operating upsets could have promoted unstable gums by permitting higher thannormal 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 crackingunit was isolated from the ethylene plant [87].

9.7 Failure scenarios and design solutionsTable 9.1 presents information on equipment failure scenarios and associated design solu-tions specific to heat transfer equipment.


9.7 Failure scenarios and design solutions 115

Figure 9.1. troubleshooting






9.8 Discussion

9.8.1 Use of Potential Design Solutions Table

To arrive at the optimal design solution for a given application, use Tables 9.1-9.4 in con-junction with the design basis selection methodology presented earlier. Use of the designsolutions presented in the table should be combined with sound engineering judgment andconsideration of all relevant factors.


9.8 Discussion 117

9.8.2 Special Considerations

This section contains additional information on selected design solutions. The informationis 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 mechanicalstresses of heat exchanger internals. The leak/rupture of tubes leads to contamination oroverpressure of the low-pressure side. Failure to maintain separation between heat transferand process fluids may lead to violent reaction in the heat transfer equipment or in thedownstream processing equipment. To make the heat transfer process inherently safer,designers must look at possible interactions between heating/cooling fluids and processfluids.

For relatively low-pressure equipment (<1000 psig), a complete failure of tubes may notbe a credible overpressure scenario if the design pressure of the low-pressure side andassociated equipment is greater than two-thirds of the design pressure of the high- pressureside [2], or if the geometry of the tube layout is such that a complete break is not physicallypossible. For high-pressure equipment (> 1000 psig), however, a complete failure shouldbe considered credible, regardless of pressure differential.

Double tube sheets or seal welding may be used for heat exchangers handling toxic chem-icals. For heat transfer problems involving highly reactive/ hazardous materials, a triple-wall heat exchanger may be used. This type of heat exchanger consists of three chambersand uses a neutral material to transfer heat between two highly reactive fluids. Alter-natively two heat exchangers can be used with circulation of the neutral fluid betweenthem.

There are known cases of cooling tower fires that have resulted from contamination ofcooling water with hydrocarbons attributable to tube leakage. Gas detectors and separa-tors may be installed on the cooling water return lines, or in the cooling tower exhaust(air) stream.

Thermal stresses can be reduced by limiting the temperature differences be-tween the inlet and outlet streams. In addition, alternate flow arrangements may beused to avoid high thermal stresses. Thermal cycling of heat transfer equipment shouldbe 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 mayreduce 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 ofreactive materials, loss of tube strength, or excessive differential thermal expansion.

Accumulation of noncondensable gases can result in loss of heat transfer capability. Heatexchangers in condensing service may need a vent nozzle, or other means of removingnoncondensable gases from the system.

External Fire (9)

Emergency relief devices are often sized for external fire. Heat transfer equipment, suchas air coolers, present a unique challenge when it comes to sizing relief devices. Theseexchangers are designed with large heat transfer areas. This large surface area may resultin very large heat input in case of external fire. Indeed, it may not be practical to installa 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 slopeddrainage, may be used to reduce the likelihood and magnitude of external fire impingingon the heat exchanger. Alternative heat exchanger designs may also be used to reducethe surface area presented to an external fire.

9.9 Troubleshooting Examples

9.9.1 Shell side temperature uncontrolled

55 Co

Controlvlave

30 Co

Organic

55-62 Co

uncontrolled

125 Co

Water

67 CoControlvlave

30 Co

Organic

55-62 Co

controlled

70 Co

Water

Bypass

Symptom: Shellside outlettemperaturee cannot becontrolled within desiredrange (55-62 C) byo

controlling flow of 125 Co

water to tubes. The heatexchanger is 4 tube pass.

Diagnosis: Heat exchanger isconsiderablyo versized for theduty (because of an alternativeservice). Temperature correctionfactors F for LMTD fluctuatewidely with small changes intube side flow

Cure: Tube side watertemperature reduced to 70oCand control valve removed.Control valve is installedin new shellside bypassline

Figure 9.5. Shell side temperature uncontrolled

9.9.2 Shell assumed banana-shape

Symptom: Shell assumedbanana shape and pipingconnections leaked. leakagebetween tube and shell side

Diagnosis: vertically cut baffleand inlets and outlets of top shellside, caused stratification ofgases at top of shell. Poordistribution of hot gases leadto unequal expansionof tubes

Cure: increase the number of bafflesfrom two to three; weld baffles in theshell; install sealing strips at edges ofbundle; installed three concentric conesin tube side inlet; install vapor belt - forshellside inlet nozzle; change bafflesfrom vertical to horizontal cut.

belows joint

487 Co

200 Co560 Co

600 Co

Figure 9.6. Shell assumed banana-shape


9.9 Troubleshooting Examples 119

9.9.3 Steam condenser performing below design capacity

Symptom: Air cooled steamcondensor performing belowdesign capacity.

Diagnosis: Careful measurementtube levels discloded that tubessloped 1/4 inch in wrong direction(rising toward condensate end)

Cure: Raise inlet end to obtain 2 inch slopetoward condensate outlet

Steam

Vent

ondensate

Figure 9.7. Steam condenser performing below design capacity

9.9.4 Steam heat exchanger flooded

When a heat exchanger ”stalls,” condensate floods the steam space and causes a varietyof 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 tothe process is greatly reduced. The control valve opens wide enough to allow flowinto the exchanger. As condensate drains out, the steam space is now greater andthe 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 tubesthat carry the process fluid. When this sub-cooled condensate is suddenly replacedby hot steam due to poor steam trap operations, the expansion and contraction ofthe tubes stress the tube joints. Constantly repeating this cycle causes prematurefailure.

• Corrosion from:

1. Flooding - A flooded heat exchanger will permit the oxygen to dissolve, as wellas 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 morereadily dissolved. Together the cool condensate and dissolved gases are ex-tremely corrosive and will tend to decrease the efficiency of the heat exchangerand reduce the heat transfer through the tubes.

2. Steam collapse - Under very low loads with the steam valve closed, the steamvolume collapses to smaller volume condensate, inducing a vacuum. Whenthe vacuum breaker opens, atmospheric air and condensate mix inside theexchanger, increasing the possibility of corrosion of the tubes, shells, tubesheet and tube supports.

3. Freezing - Steam/air coils cannot afford poor condensate drainage, especiallyif the coil experiences air below freezing temperature. Condensate backed upinside the coil will freeze, often within seconds, depending on the air temper-ature. A low temperature detection thermostat is recommended on the coilleaving side to sense freezing conditions. As we previously explained, the onlyway to avoid ”stall” is to eliminate back pressure on the steam trap. There area number of options available for designing a system that greatly reduces therisk of ”stall.” The following are two such options:

• Install the heat exchanger in a position so that the condensate freely drains bygravity to the condensate return line. In many cases this is not possible becauseof existing piping around the area in which the heat exchanger is needed (e.g., theheat exchanger is installed at a level lower than the condensate return tank).

• Use an electric or pressure driven condensate pump package installed below thesteam trap to pump condensate back to the boiler.

In actual practice, the first option may not be possible, and so the use of electric orpressure driven pumps to return condensate to the boiler room should be considered.

10 Unresolved problems in the heat exchangers de-

sign1. 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 trend1. 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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131

A Heat transfer coefficient

A.1 Single phase

A.1.1 Inside tube: Turbulent flow

Nu = CReaPrb

(µ

µw

)c

, (A.1)

whereNu = hde

kNusselt number

Pr = Cpµk

Prandtl number

Re ρudµ

Reynolds number

de4AP

hydraulic diameterA cross-sectional areaP wetted perimeteru fluid velocityµw fluid viscosity at the tube wall temperaturek fluid thermal conductivityCp fluid specific heat

C =

0.021 gases0.023 non-viscous liquid0.027 viscous liquid

a = 0.8b = 0.3 for coolingb = 0.4 for heatingc = 0.14

A.1.2 Inside tube: Laminar flow

Nu = 1.86

(RePr

d

L

)1/3 (µ

µw

)0.14

, (A.2)

A.1.3 Shell side

For the shell side heat transfer coefficient there are a number of methods the include:

• Kern’s method

• Donohue’s method

• Bell-Delaware method

• Tinker’s method

Besides these methods there is some proprietary methods putout by various organizationfor use by their member companies. A number of these method are based on one of theabove methods. Some are based upon a judicious combination of methods 3 and 4 aboveand supplemented by further research data. Among the most popular of the proprietarymethods, judged by their large clientele are


132 A Heat transfer coefficient

• Heat Transfer Research Inc. (HTRI), Alliambra, california. This method is alsoknown 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 inCoulson and Richardson’s

Nu = 0.36Re0.55Pr1/3

(µ

µw

)0.14

, (A.3)

whereNu = hde

kNusselt number

Pr = Cpµk

Prandtl numberRe = Gde

µReynolds number

de = 4AP

hydraulic diameterA = cross-sectional flow areaP = wetted perimeterG = M

AsMass flux

As = (pt−do)DslBpt

fluid viscosity at the tube wall temperature

pt = pitch diameterDs = shell diameterlB = Baffle spacing

Hydraulic diameter (Fig. A.1)

de =

p2t−πd2

o/4

πdofor square pitch

0.87p2t /2−πd2

o/8

πdo/2for equilateral triangular pitch

p t

do

Square pitch

p t

Equilateral triangular pitch

As

Cross-flow area

Figure A.1. Tube arrangement


A.2 Condensation 133

A.1.4 Plate heat exchanger

Nu = 0.26Re0.65Pr0.4

(µ

µw

)0.14

, (A.4)

where

Nu = hde

k= Nusselt number

Pr = Cpµk

= Prandtl number

Re = ρupde

µ= Gde

µ= Reynolds number

de = hydraulic diameter, taken as twice the gap between the platesA = cross-sectional flow areaP = wetted perimeterG = M

Af= Mass flux

Af = cross-sectional area for flowup = channel velocityM = mass flow rate

A.2 Condensation

A.2.1 Condensation on vertical plate or outside vertical tube

hm = 0.943

(k3ρ∆ρgλ

µ∆TL

)1/4

, (A.5)

where

hm = mean heat transfer coefficientL = lenth of the plate or the vetical tubek thermal conductivity of the saturated liquid filmρ = liquid densityµ = liquid viscosityλ = latent heat of evaporization∆T = Ts − Tw temperature difference across the condensate filmg = acceleration due to gravityTs = saturation temperature of the condensate filmTw = wall temperature

A.2.2 Condensation on external horizontal tube

hm = 0.725

(k3ρ∆ρgλ

µ∆Tdo

)1/4

, (A.6)

wheredo = out side diamter of the tube

A.2.3 Condensation on banks of horizontal tube

hm = 0.725

(k3ρ∆ρgλ

µ∆TJdo

)1/4

, (A.7)



whereJ = number of tubes in a row (Fig. ??)

In the above equation the condensate film properties save the latent heat of vaporizationare evaluated at the film temperature.

Tf =Ts + Tw

2, (A.8)

the latent heat of vaporization is evaluated at the condensate temperature. For the caseof subcooling or superheating the heat transfer coefficient is corrected by substituting thecorrected latent heat the heat transfer equation (Rohsenow et al. [121] and Carey [18]) inNusselt [109])

λ∗ = λ + 0.68cp∆T . (A.9)

A.2.4 Condensation inside horizontal tube

hm = 0.555

(k3ρ∆ρgλ

µ∆Td

)1/4

, (A.10)

A.3 Two phase flow: Pure fluid

A.3.1 Steiner [140] correlation

Steiner [140] has considered the two phase heat transfer coefficient h as a combination ofthe convective and the nucleate part using an asymptotic model as:

h =(h3

n + h3c

)1/3, (A.11)

where hn and hc is the nucleate and convective boiling heat transfer coefficient respectively.The convective boiling heat transfer coefficient for a completely wetted tube (i.e. all typesof flow patterns save stratified and stratified-wavy flow) is calculated as

hc

hL0

=

(1− x) + 1.2x0.4(1− x)0.01

(ρL

ρG

)0.37 +

hG0

hL0

x0.01

1 + 8(1− x)0.7

(ρL

ρG

)0.67

−2

−0.5

. (A.12)

The heat transfer coefficients hL0 and hG0 are those of single phase flow, assuming thatthe total mass velocity is pure liquid or pure vapor respectively. They are calculated inthe case of a fully developed turbulent flow from the Gnielinski [46] model

Nu =(ξ/8)(Re− 1000)Pr

1 + 12.7(ξ/8)0.5(Pr2/3 − 1), (A.13)

taken in to account the respective dimensionless group NuL0, NuG0, ReL0, ReG0, Prl andPrg. These dimensionless groups are defined as

NuL0/G0 =hL0/G0d

kL/G

, (A.14)


A.3 Two phase flow: Pure fluid 135

ReL0/G0 =md

µL/G

, (A.15)

PrL/G =µL/Gcp,L/G

kL/G

. (A.16)

The friction factor isξ = (1.82logRe− 1.62)−2. (A.17)

For a partial wetting of the tube (stratified or stratified-wavy flow) the average heattransfer coefficient at the tube circumference under the thermal boundary condition of aconstant 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 thetube 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 thevapor-occupied part of the tube cross-section

dh = d

(ϕ− sin ϕ

d + 2 sin(ϕ/2)

), (A.19)

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

ReG =mxdhyd

εµG

, (A.20)

and

hG =NuGkG

dhyd

. (A.21)

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

ε =x

ρG

(1 + 0.12(1− x))

(x

ρG

+1− x

ρL

)+

1.18(1− x)[gσ(ρL − ρG)]1/4

mρ .1/2L

−1

(A.22)

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

ε =fG

fG + fL

. (A.23)

With some mathematical manipulation of equation A.23 the non-wetted perimeter cancalculated 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) inthe vapor phase, the scaling parameter Φ of equation A.18 can thus be calculated as

Φ =ϕG

2π, (A.25)

where ϕG = 0.5ϕ.



ϕ

d

hfL

fG

Ui

UG

UL

Figure A.2. Cross-section and perimeter parts of the vapor flow in a horizontal tube.

The local nucleate boiling heat transfer coefficient hnb of a horizontal tube is estimatedas

hnb

ho

= ψCf

(q

qo

)n(pr)

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

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

The pressure function is given as

F (pr) = 2.692p0.43r +

(1.6p6.5

r

1− p4.4r

), (A.27)

and the mass flux function is given as

F (m, x) =m

mo

0.251− p0.1

r

(q

qcr,nb

)0.3

x

, (A.28)

whereqcr,cb = 2.79qcr,0,1p

0.4r (1− pr) . (A.29)

The critical value of qcr,0,1 at a reduced pressure pr of 0.1 is given as

qcr,0.1 = 0.13∆hV,0ρ0.5G,0[σog(ρL,0 − ρG,0)]

0.25 . (A.30)

The function for the effect of surface roughness and tube diameter is F (Ra) =(Ra/Rao)0.133

and F (d)=(do/d)0.5 respectively. The pressure dependence of the heat flux exponent n(pr)can be predicted as

n(pr) = 0.9− 0.3p0.3r . (A.31)

The experimental value of the specific constant Cf for a number of substances is be foundin VDI-Warmeatlas[157], for example for water it is 0.72. In absence of an experimentalvalue 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 stratifiedand 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-Warmeatlas[157]).

Table A.1 shows the reference factors for the nucleate boiling heat transfer coefficient forR134a and R290.

Table A.1. Values of the reference parameters used in evaluation of the local nucleate boilingheat transfer coefficient.

Refrigerant ho qo Rao do

W/m2K W/m2 m m

R134a 3,500 20,000 10−6 0.01R290 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 twophase heat transfer coefficient is

h =ϕdryhG + (2π − ϕdry)hwet

2π. (A.33)

The vapor heat transfer coefficient hG is determined by using the Dittus-Boelter [33]correlation as

hG = 0.023Re0.8G P 0.4

rG

kG

d, (A.34)

with Reynold number given as

ReG =mxd

εµG

, (A.35)

where ε is the void fraction given by the Rauhani [117] model (equation A.22). The heattransfer coefficient on the wetted portion of the tube is

hwet = 3

√h3

n + h3c . (A.36)

The nucleate boiling heat transfer coefficient hn is given by the Cooper [27] model as

hn = 55p0.12r (−0.4343 ln pr)

−0.55M−05q . (A.37)

The convective heat transfer coefficient is given by a modified form of the Dittus-Boelter[33] model as

hc = 0.0133Re0.69L P 0.4

rL

kL

d. (A.38)

The liquid Reynolds number is given as

ReL =4m(1− x)δ

(1− ε)µG

. (A.39)

where δ is the liquid film thickness it is given as

δ =πd(1− ε)

2(2π − ϕdry), (A.40)



where ϕdry is

ϕdry = ϕstrat(mwavy − m)

(mwavy − mstrat), (A.41)

where ϕstrat is calculated iteratively from equation A.24. The mass flux under a stratifiedand wavy flow pattern is

mstrat =(226.3)2fLf 2

GρG(ρL − ρG)µLg cos Θ

0.3164(1− x)1.75π2µ0.25L

, (A.42)

and

mwavy =

16f 3

GgdρLρG

x2π2(1− (2hL − 1)2)0.5

[π2

25h2L

(1− x)F1(q) ×(

We

Fr

)F2(q)

L+

1

cos Θ

]0.5

+ 50 ,

(A.43)respectively. The parameters fL, fG, hL are defined in Fig.A.2. Θ is the angle of inclinationto the horizontal and

F1(q) = 646.0

(q

qcrit

)2

+ 64.8

(q

qcrit

); F2(q) = 18.8

(q

qcrit

)+ 1.023 . (A.44)

The stratified-wavy flow model is also valid for the stratified flow patten with ϕstrat

replacing ϕdry and for the annular flow condition with ϕdry is set to zero and the filmthickness δ is set to (1− ε)d/4.

A.3.3 Kandlikar [70] correlation

The flow boiling heat transfer coefficient for a pure fluid is given by Kandlikar [70] as

h = max(hn, hc) , (A.45)

wher the subscript n and c in equation A.45 refers to the nucleate and convective boilingrespectively. The convective and the nucleate boiling part is given as

hn = 0.6683Co−0.2(1− x)0.8hL0f(FrL0) + 1058.0Bo0.7(1− x)0.8FFlhL0 , (A.46)

and

hc = 1.136Co−0.9(1− x)0.8hL0f(FrL0) + 667.2Bo0.7(1− x)0.8FFlhL0 , (A.47)

respectively, where FrL0 is the liquid Froude number, Bo is the boiling number and Cois the convection number. These dimensionless groups are defined as

FrL0 =m

ρLgd, Bo =

q

m∆hv

, Co =

(ρG

ρL

)0.5 (1− x

x

)0.8

. (A.48)

The function f(FrL0) is defined as

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

f(FrL0) = 1 FrL0 ≥ 0.04 ,

where FFl is a fluid-surface parameter related to the nucleation characteristic. For alltype of fluids flowing in a stainless tube it is taken as 1. The single phase heat transfer



coefficient hL0 is obtained from the Petukhov and Popov [114] correlation or Gnielinski[46] correlation. The Petukhov and Popov [114] correlation is valid in the range of 0.5 ≤PrL ≤ 2000 and 104 ≤ ReL0 ≤ 5× 106 and it is given as

NuL0 =hL0d

k=

ReL0PrL(ξ/2)

1.07 + 12.7(P2/3rL − 1)(ξ/2)0.5

. (A.49)

The Gnielinski [46] correlation (equation A.13) is valid in the range of 0.5 ≤ PrL ≤2000 and 2300 ≤ ReL0 ≤ 5 × 104. The friction factor ξ in equation A.49 is given byequation A.17.

A.3.4 Chen [19] correlation

Chen [19] postulated that the heat transfer coefficient is made of two parts: a) a micro-convective (or nucleate boiling) portion hn and b) a macro-convective (or forced convec-tive) portion hc as

h = hcF + hnS , (A.50)

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

hc = 0.023kL

dRe0.8

L Pr0.4L , (A.51)

where

ReL =(1− x)md

µL

, P rL =cpLµL

kL

, (A.52)

The suppression factor for the convection part is

F =

1 if 1/Xtt > 0.1

2.35[

1Xtt

+ 0.213]0.736

if 1/Xtt ≤ 0.1

,

and the Martinelli parameter Xtt is given as

X =(

1− x

x

)0.875(

ρG

ρL

)0.5 (µL

µG

)0.125

. (A.53)

The nucleate boiling heat transfer coefficient is

hn = 0.00122k0.79

L c0.45p,L ρ0.49

L

σ0.5µ0.29L ρ0.24

G ∆h0.24V

∆T 0.24sat ∆p0.75

sat , (A.54)

where

∆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× 10−6Retp

. (A.56)



A.3.5 Gungor and Winterton [52] correlation

The Gungor and Winterton [52] correlation is a modified form of the Chen [19] correlationgiven by equation A.50 with the nucleate boiling calculated from the Cooper [27] corre-lation given by equation A.37. The suppression factor for the convection part is definedas

F =

(1 + 24, 000Bo1.16 + 1.37(1/xtt)0.86)Fr

(0.1−2FrL)L if Fr < 0.05

1 + 24, 000Bo1.16 + 1.37(1/xtt)0.86 if Fr ≥ 0.05

,

and the suppression factor for the nucleate part is

S =

(1 + 0.00000115F 2ReL)−1Fr1/2L if Fr < 0.05

(1 + 0.00000115F 2ReL)−1 if Fr ≥ 0.05

,

The convective boiling part is calculated from the Dittus-Boelter [33] correlation (equationA.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 boilingrespectively. The convective heat transfer coefficient is defined as

hc = 1.8hLN−0.8 , (A.58)

where

N =

Co FrL > 0.04

0.38Fr−0.4L Co FrL < 0.04

,

where hL is calculated using the Dittus-Boelter [33] correlation (equation A.51). Thenucleate boiling heat transfer coefficient is calculated as follows

• For N > 1

hn =

230hLBo0.5 Bo > 0.0003

1 + 46hLBo0.5 Bo < 0.0003.

• For 1 > N > 0.1hn = FhLBo0.5 exp(2.74N−0.1) . (A.59)

• For N < 0.1hn = FhLBo0.5 exp(2.47N−0.15) , (A.60)

where

F =

14.7 Bo > 0.0011

15.43 Bo < 0.0011.



A.3.7 Schrock and Grossman [129] correlation

A very simple correlation is given by Schrock and Grossman [129] as

h = 1.91hL

[104 ×Bo + 1.5

(1

Xtt

)2/3]0.6

, (A.61)

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

A.3.8 Dembi et al. [30] correlation

The Dembi et al. [30] correlation is based on the asymptotic model given by equationA.11 with the nucleate and convection part given as

hn = 23388.5kL

d

(q

ρG∆hV $

)0.64 (gd

∆hV

)0.27 (m2d

ρL∆hV $

)0.14

, (A.62)

and

hc = 0.115kL

d

[x4(1− x)2

]0.11(

m2∆hV

ρLgσ

)0.14

P 0.27rL , (A.63)

respectively. The parameter $ is defined as

$ = 0.36× 10−3p−1.4r . (A.64)

A.3.9 Klimenko [84] correlation

The Klimenko [84] correlation is based on the asymptotic model given by equation A.11with the convection part given by the Dittus-Boelter [33] correlation equation A.51 andthe nucleate boiling is

hn =

hn1 NCB < 1.6× 104

hn2 NCB > 1.6× 104 ,

where

hn1 = 7.4× 10−3

(kw

kL

)0.15

Pe0.6K0.5p Pr

−1/3L , (A.65)

hn2 = 0.087kL

b

(kw

kL

)0.09

Re0.6m

(ρG

ρL

)0.2

Pr1/6L , (A.66)

Pe =

(qb

∆hV ρGaL

), Kp =

p√σg(ρL − ρG)

, b =

√2σ

g(ρL − ρG), (A.67)

Rem =wmb

νL

, wm =m

ρL

[1 + x

(ρL

ρG

− 1

)], Re∗ =

qb

∆hV ρGνL

, NCB =Rem

Re∗

(ρL

ρG

).

(A.68)



A.3.10 Jung et al. [64] correlation

The Jung et al. [64] correlation is a modified form of the Chen [19] correlation. Theconvection 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 inVDI-Warmeatlas [157] correlation as

hn = 207kL

b.d

[q(b.d)

kLTs

]0.745 (ρG

ρL

)0.581

P 0.533rL , (A.69)

where

(b.d) = 0.511

(2σ

g(ρL − ρG)

)0.5

, (A.70)

F = 2.37(0.29 +

1

Xtt

), (A.71)

S =

4048X1.22tt Bo1.13 Xtt < 1

2.0− 0.1X−0.28tt Bo−0.33 1 ≤ Xtt ≤ 5

.

A.4 Two phase flow: Mixture

A.4.1 Steiner [140] correlation

Steiner [140] has extended his pure component asymptotic model to mixture. The nucleatepart of the heat transfer coefficient is suppressed using the Schlunder [126] suppressionfactor for the nucleate boiling. The Schlunder [126] suppression factor is based on theheat 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 andk stands for the more volatile and less volatile component respectively. βL/B0 = 5× 105

is the mass transfer coefficient. The ideal nucleate boiling heat transfer coefficient for amixture hid,n is calculated from the heat transfer coefficient of pure components as

hid,n =

[∑ xi

hi,n

]−1

, (A.73)

and Bo/βL = 5.103 and ρL and ∆hV is the ideal density and enthalpy of evaporation ofthe mixture respectively. x and y is the liquid and vapor mole fraction of the more volatilecomponent respectively.

The same approach applies also to the convective part for the liquid-liquid immisciblemixture. That is to say for a liquid-liquid miscible mixture the convective suppressionfactor 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 mixturesas

• 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 respectivelyusing the mixture properties.

• Region II: Moderate diffusion-induced suppression region

h = hc , (A.76)

where hc is given by equation A.77 with the properties of the mixture.

• Region III: Severe diffusion-induced suppression region: 0.03< V1 < 0.2 and Bo ≤1E−4; V1 ≥ 0.2

h = 1.136Co−0.9(1− x)0.8hL0f(FrL0) + 667.2Bo0.7(1− x)0.8FFlhL0FD , (A.77)

where

V1 =

[(cpL

∆hV

) (a

D12

)0.5

|y − x|(

dT

dx

)], (A.78)

FD =0.678

1 + V1

. (A.79)

A.4.3 Bennett and Chen [8] correlation

Bennett and Chen [8] has extended the Chen [19] correlation (equation A.50) for mix-ture. Here both the convective and the nucleate parts are suppressed. The convectionpart which is calculated for the original Chen [19] correlation with mixture properties issuppressed using the following suppression factor

Fc =Tw − Tph

Tw − Ts

, (A.80)

where Tw, Tph, and Ts is the wall, equilibrium temperature and saturation temperaturerespectively. The nucleate part is also calculated using the original Chen [19] model forthe pure substance with mixture properties. It suppressed using the the suppression factorgiven by equation A.79.

A.4.4 Palen [111] correlation

Palen [111] has extended the original Chen [19] correlation for pure component (equationA.50) to mixture similar to the Bennett and Chen [8] correlation. However, only thenucleate part is suppressed using the following suppression factor

Fd = exp(−0.027∆Tbp) , (A.81)

where ∆Tbp is difference between the dew and bubble point temperature of the mixture.



A.4.5 Jung et al. [64] correlation

Jung et al. [64] have extended their pure substance correlation to the mixture. The nu-cleate boiling heat transfer coefficient is replaced by the ideal one given by equation A.73.The convective part is suppressed using the following suppression factor

Fc = 1.0− 0− 35|y1 − x1|1.56 . (A.82)

For the nucleate part the following suppression factor is employed

Fn =1

[1 + (b2 + b3)(1 + b4)](1 + b5)2 , (A.83)

where

b2 = (1− x1) ln

(1.01− x1

1.01− y1

)+ x1 ln

(x1

y1

)+ |y1 − x1|1.5 , (A.84)

b3 =

0 x1 ≥ 0.01

(x1

y1

)0.1 − 1 x1 < 0.01

,

b4 = 152

(p

pc,1

)0.66

, (A.85)

b5 = 0.92|y1 − x1|0.001

(p

pc,1

)0.66

, (A.86)

andx1

y1= 1 for x1 = y1 = 0 ,

x1 and y1 is the liquid and vapor mole fraction of the more volatile component respec-tively. p and pc,1 is system pressure and critical pressure of the more volatile componentrespectively.


145

B Pressure drop

B.1 Single phaseThe pressure drop due to friction exists because of the shear stress between the fluid andthe tube wall. Estimation of the friction pressure drop is somewhat more complex andvarious approaches have been taken, for example the frictional pressure gradient is givenas

−(

dp

dz

)

f

=4τo

d=

4fm2

2dρ, (B.1)

where m is the mass flux in kg/m2s and f is the friction factor calculated using a Blasius-type model as

f =

0.3164Re0.25 Re ≥ 2320

64Re

Re < 2320 .

Integration of equation B.1 yields

∆p =4fm2

2ρ

L

d, (B.2)

B.2 Two phaseIn flow boiling, the temperature drops in the direction of flow as a result of the pressuredrop. This results in a change in the driving force (temperature difference) for the heattransfer along the flow path. Thus beside the heat transfer coefficient, knowledge of thepressure drop is of paramount importance in the design of the evaporator. In the presentwork the pressure drop is measured simultaneously with the heat transfer coefficient alongthe test section.

The momentum balance implies that the two phase pressure gradient is composed of threecomponents 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 andhydrostatic pressure gradient respectively. For a horizontal tube the hydrostatic pressuregradient diminishes. The acceleration pressure drop is caused by the change in momentumin both the liquid and vapor phases. The change in the momentum stems from the changein 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 dropdiminishes 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 staticpressure drop due to acceleration. The most widely accepted models include homogenousor separated flow models. The separated flow model is also widely known as the het-erogenous model. In the homogenous model the static pressure drop due to accelerationis

−(

dp

dz

)

a

= m2 d

dz

[x

(1

ρL

− 1

ρG

)+

1

ρL

]. (B.4)

The energy balance in a small unit length dz along the test tube yields

dx

dz=

4q

m∆hvd. (B.5)


146 B Pressure drop

Substitution of equation B.5 into equation B.4 yields the pressure drop due to accelerationas

∆pa =4qm

d∆hvρG

(1− ρG

ρL

)∆L . (B.6)

In the separated flow model the static pressure drop due to acceleration can be derivedfrom the momentum balance as

−(

dp

dz

)

a

= m2 d

dz

[x2

ερG

+(1− x)2

(1− ε)ρL

]. (B.7)

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

−∆pa = −(po − pi)a = m2

[x2

2

εoρG,o

+(1− xo)

2

(1− εo)ρL,o

− x2i

εiρG,i

− (1− xi)2

(1− εi)ρL,i

]. (B.8)

The void fraction ε may be obtained using the Rauhani [117] model which is given as:

ε =x

ρG

(1 + 0.12(1− x))

(x

ρG

+1− x

ρL

)+

1.18(1− x)[gσ(ρL − ρG)]1/4

mρ1/2L

−1

, (B.9)

where ρL and ρG is the liquid and vapor density respectively, which are calculated from thefundamental equation of state of Tillner-Roth and Baehr [152] for R134a. g is accelerationdue to gravity, σ is the surface tension, m is the mass flux and x is the quality. The surfacetension is calculated using the method of Lucus [92] given in VDI-Warmeatlas [157].

The pressure drop due to friction exists because of the shear stress between the fluid andthe tube wall. Estimation of the friction pressure drop is somewhat more complex andvarious 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=

2ξm2

dρH

, (B.10)

where ξ is the two phase friction factor calculated by a Blasius-type model as

ξ =

0.3164Re0.25 Re ≥ 2320

64Re

Re < 2320 .

and the homogenous densityρH is given as

1

ρH

=1− x

ρL

+x

ρG

. (B.11)

The two phase Reynolds number Re is

Re =md

ηTP

, (B.12)

where ηTP is a two-phase viscosity. A variety of methods have been proposed to calculatethe two phase viscosity, a commonly used one being that proposed by McAdams et al. [95]

1

ηTP

=1− x

ηL

+x

ηG

, (B.13)


B.2 Two phase 147

where ηL and ηG are the liquid and vapor viscosity.

In the separated flow model the two phase frictional pressure drop is related to that forsingle phase as (

dp

dz

)

f

=

(dP

dz

)

f,L/G

ΨG/L , (B.14)

where Ψ is the two phase multiplier. There exist a number of correlations for the predictionof 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 theprediction of the two phase multiplier Ψ of the separated flow model. These models arepresented in the following subsections.

B.2.1 Friedel [42] model

ΨL0 = E +3.24FH

Fr0.045We0.035, (B.15)

where

E = (1− x)2 + x2ρLfG0

ρGfL0

, (B.16)

F = x0.78(1− x)0.24 , (B.17)

H =

(ρL

ρG

)0.91 (µG

µL

)0.19 (1− µG

µL

)0.7

, (B.18)

Fr =m2

gdρ2H

, (B.19)

We =m2d

σρH

, (B.20)

d is tube diameter, σ is the surface tension and %H is the homogenous density given byequation B.11. fG0 and fL0 are the friction factors defined by a Blasius-type model as

fL0/G0 =0.079

Re0.25L0/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

ψ2L = 1 +

C

X+

1

X2, (B.22)

ψ2G = 1 + C.X + X2 , (B.23)

where X is the Martinelli parameter and the value of the coefficient C is given in Table B.1.

The range of the applicability of the Lockhart and Martinelli [91] correlation is µL/µG >1000and m <100 kg/m2s.


148 B Pressure drop

Table B.1. Value of C for the Lockhart and Martinelli [91] correlation.Liquid Gas Subscript C

Turbulent Turbulent tt 20Viscous Turbulent vt 12

Turbulent Viscous tv 10Viscous Viscous vv 05

B.2.3 Chisholm [22] model

In the Chisholm [22] model the two phase friction multiplier is

ΨL0 = 1 + (Y 2 − 1)[Bx(2/n−1)(1− x)(2/n−1) + x1−n

], (B.24)

where

Y 2 =(dpf/dz)G0

(dpf/dz)L0

, (B.25)

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

B =55

m1/20 < Y < 9.5 , (B.26)

B =520

Y m1/29.5 < Y < 28 , (B.27)

B =15000

Y 2m1/228 < Y . (B.28)

The range of the validity of the Chisholm [22] correlation is µL/µG > 1000 and m > 100kg/m2s.


149

C Physical propertiesThe fluid physical properties required for heat exchanger design are divided in thermo-dynamic and trasport properties. The transport properties include viscosity, thermalconductivity, surface tension and diffusion coefficient are generally calculated from theexisting correlations (Pery and Coulson). The thermodynamic properties include dem-sity, specific heat temperature, pressure (vapor), enthalpy, latent heat of evaporation.Beside the fluid properties the thermal conductivity of the material is necessary for theevaluation of heat transfer coefficient. The thermodynamic properties are evaluated usingcritical tables.

C.1 Physical properties: Pure fluid

C.1.1 Specific heat

The specific heat of the ideal gas is given in as

Cp = CPV APA + (CPV APB)T + (CPV APC)T 2 + (CPV APD)T 3 (C.1)

Where T is in K and CPVAPA, CPVAPB, CPVAPC, CPVAPD are constant in idealgas heat capacity. These constant are given in Appendix A for organic and inorganiccompounds.

C.1.2 Vapor pressure

The vapor pressure is generally predicted using Antonie equation as

ln p = ANTA− ANTB

T + ANTC(C.2)

where T is in K and ANTA, ANTB,ANTC are Anonie equation constant. These constantare given in Appendix D for organic and inorganic compounds.

C.1.3 Liquid viscosity

The liquid viscosity is given as:

log µ = V ISA(

1

T− 1

V ISB

)(C.3)

where VISA, VISB are constants in the liquid viscosity equation. These constant aregiven in Appendix D for organic and inorganic compounds.

C.1.4 Vapor dynamic viscosity VDI-Warmeatlas [157]

Lucas and Luckas [92] in VDI-Warmeatlas [157] have recommended the following proce-dure for the calculation of the vapor viscosity.

η = (ηξ)rFpFQ1

ξ, (C.4)

for Tr ≤ 1 and pr ≤ ps/pc

(ηξ)r = 0.600 + 0.760pαr + (6.990pβ

r − 0.6)(1− Tr) , (C.5)

withα = 3.262 + 14.98p5.508

r and β = 1.390 + 5.746pr , (C.6)


150 C Physical properties

for 1≤ Tr ≤ 40 and 0≤ pr ≤ 100

(ηξ)r = (ηoξ)

[1 +

ApEr

BpFr + (1 + CpD

r )−1

], (C.7)

where ηo is the low pressure viscosity given as

ηoξ = [0.807T 0.618r − 0.357 exp(−0.449Tr) + 0.340 exp(−4.058Tr) + 0.018]F o

p F oQ , (C.8)

and ξ is given as

ξ =[Tc]

1/6[R]1/6[Na]1/3

[M ]1/2[pc]2/3, (C.9)

where Na is the Avagadro number in kmol. The coefficients of equation C.7 are given as

A =a1

Tr

exp(a2Tγr ) , (C.10)

B = A(b1Tr − b2) , (C.11)

C =c1

Tr

exp(c2Tδr ) , (C.12)

D =d1

Tr

exp(d2Tεr ) , (C.13)

E = 1.3088 , (C.14)F = f1 exp(f2T

ςr ) . (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 -0.3286b1 1.6553 b2 1.2723 d1 1.7368 d2 2.2310 δ -37.7332f1 0.9425 f2 −0.1853 ς 0.4489 ε -7.6351

Fp = 1 + (F op − 1)

[(ηξ)r

ηoξ

]−3

, (C.16)

and

FQ = 1 + (F oQ − 1)

[(ηξ)r

ηoξ

]−1

− 0.007

[ln

((ηξ)r

ηoξ

)]4

, (C.17)

where F op and F o

Q is low-pressure polarity and quantum factors respectively. These factorsare

F op = 1 , 0 ≤ µr < 0.022 , (C.18)

F op = 1 + 30.55(0.292− Zc)

1.7 , 0.022 ≤ µr < 0.075 , (C.19)

F op = 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 oQ = 1.0 for all substances other than

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

µr =µ2pc

(kTc)2, (C.21)

where the dipole moment µ for the gases is given in VDI-Warmeatlas [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 Fenghouret al. [40] is

η = ηo(T ) + η1(T )ρ + η2(ρ, T ) , (C.22)

The first term of the expansion is the dilute gas term which is given as

ηo(T ) = 100[0.021357

0.29572

](MT )1/2

exp(Ω), (C.23)

where M is the molecular weight in g/mol, T is the temperature in K. The collisionintegral Ω is defined as

Ω(T ) =

C(1) + C(2) log

(kT

ε

)+

4∑

n=3

C(n)

[log

(kT

ε

)]n, (C.24)

where ε/k=386 K and the value of the coefficient C is given in table C.2.

Table C.2. Coefficients for the Collision integral Ω (equation C.24).

C(1) 4.9931822 C(2) -0.61122364 C(3) 0.18535124 C(4) -0.1116094

The second term of equation C.22 represents the contribution of the moderately densefluid

η1(T ) = Fv(T )ηo(T )ρ , (C.25)

where

Fv(T ) = C

A(1) +13∑

i=2

A(i)

[log

(kT

ε

)]−(i−1)2

, (C.26)

where C=0.6022137/0.29573 and the value of the coefficient A is given in table C.3

Table C.3. Coefficients of equation C.26.i A i A

1 -0.17999496×101 2 0.466692621×102

3 -0.53460794×103 4 0.33604074×104

5 -0.13019164×105 6 0.33414230×105

7 -0.58711743×105 8 0.71426686×105

9 -0.59834012×105 10 0.33652741×105

11 -0.12027350×105 12 0.24348205×104

13 -0.120807957×103

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

η2(ρ, T ) =3∑

i=1

F (i, T )ρi+1 , (C.27)



where

F (i, T ) =

1 0.219664285(

εkT

)2 − 0.83651107× 10−1(

εkT

)4

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

εkT

)

3 0.167668649× 10−3(

εkT

)2 − 0.149710093× 10−3(

εkT

)3+

0.77012274× 10−4(

εkT

)4

The Fenghour et al. [40] correlation for the vapor viscosity of ammonia has an uncertaintyof 2% in the temperature range of T < Tc.

C.1.6 Surface tension

Lucas and Luckas [92] in VDI-Warmeatlas [157] have recommended the following corre-lation for the calculation of the surface tension

σ = p2/3c T 1/3

c

(1− Tr

a

)m

b , (C.28)

where the reduced pressure and temperature are defined as

pr =p

pc

, Tr =T

Tc

, , (C.29)

respectively.

For a polar fluid like R134a the following quantities are valid

a = 1 , (C.30)

b = 0.1574 + 0.359ω − 1.769X − 13.69X2 − 0.510ω2 + 1.298ωX , (C.31)

m = 1.210 + 0.5385ω − 14.61X − 32.07X2 − 1.656ω2 + 22, 03ωX , (C.32)X = lgpsr(Tr = 0.6) + 1.70ω + 1.552 . (C.33)

where ω is the acentric factor and it is given by Pitzer in VDI-Warmeatlas [157] as Thesurface tension given by equation C.28 is in 10−5 N/cm. Its level of uncertainty as givenby 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× 10−5Cp

(ρ

M

)1/3

. (C.34)

where k thermal conductivity W/moC, M is the molecular mass, Cp speific heat capacity

(kJ/kg oC), ρ density (kg/m3)

C.1.8 Thermal conductivity for gases

k = µ(Cp +

10.4

M

). (C.35)

where k thermal conductivity W/m oC, M is the molecular mass, Cp specific heat capacity(kJ/kg oC), µ viscosity in (mNs/m2)


C.2 Physical properties: Mixture 153

C.1.9 Specific enthalpy

For the vapor phase, the deviation of the specific enthalpy from the ideal state can beillustrated using Redlich-Kwong equation written as

z3 + z2 + z(B2 + B − A) = 0 . (C.36)

where z is the compressibilty factor defined as

z =pv

RT. (C.37)

and

A =aP

R2T 2.5, B =

bp

RT. (C.38)

h = ho + RT +∫ v

0

[(T

dP

R2T 2.5dT

)− p

]dv . (C.39)

C.2 Physical properties: Mixture

C.2.1 Liquid dynamic viscosity of mixtures

For a liquid mixture which contains one or more polar constituents Reid et al. [118]recommended the following model for the calculation of the mixture liquid viscosity

ln ηm =n∑

i=1

xi. ln ηL,i + 2.x1.x2.G12 , (C.40)

where xi is the mole fraction of the component i, ηL,i is the viscosity of the component iin 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 thermalconductivity 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 kinetictheory (Reid et al. [118]) as

ηm = ηom + ∆η , (C.41)

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

the high pressure viscosity

ηom =

n∑

i=1

yiηG,i∑nj=1 yiφij

, (C.42)

where yi is the mole fraction of the component i and ηi is the viscosity of the purecomponent i. φij is a parameter which may be estimated as

φij =

[1 + (ηG,i/ηG,j)

0.5(Mj/Mi)0.25

]2

[8(1 + Mi/Mj)]0.5, (C.43)

φji =ηG,j

ηG,i

Mj

Mi

φij . (C.44)



The high pressure correction term is estimated as

∆η =0.497.10−6

[exp(1.439ρr,m)− exp(−1.111ρ1.858

r,m )]

T1/6c,mM−0.5

m p−2/3c,m

. (C.45)

The pseudo critical properties of the mixture are calculated as

Tc,m =∑

j=1

yjTc,j, υc,m =∑

j

yjυc,j, Zc,j =pc,jυc,j

RTc,j

, Zm =∑

j

yjZc,j, (C.46)

Mm =∑

j=1

yjMj, ρc,m =Mm/1000

υc,m

, ρr,m =ρm

υc,m

, pc,m =RTc,mZc,m

υc,m

, (C.47)

where T is in K, p is in Mpa, υc,m is in m3/kmol, ρr,m is in kg/m3, M is in g/mol and ηm

is in kg/ms. The error associated with this model is seldom exceeded 3 to 4% (Perry andGreen [112]).

C.2.3 Liquid thermal conductivity of mixtures

Reid et al. [118] have recommended a Filippov-like model for the prediction of the thermalconductivity of a liquid mixture as

λm =2∑

i=1

XiλL,i − 0.72X1X2|λL,2 − λL,1| , (C.48)

where X1 and X2 is the weight fraction of the component 1 and 2 respectively and λ1 andλ2 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 therelationship 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-pressurethermal conductivity of the pure component i. For a binary mixture of two non-polargases or a non-polar and a polar gas, Aij may be calculated by the model given by Perryand Green [112] as

Aij =

[1 + (λtr,i/λtr,j)

0.5(Mj/Mi)0.25

]2

[8(1 + Mi/Mj)]0.5, (C.50)

withλtr,i

λtr,j

=Γj

Γi

exp(0.0464Tr,i)− exp(−0.2412Tr,i)

exp(0.0464Tr,j)− exp(−0.2412Tr,j), (C.51)

where M is the molecular weight and Γ is defined as

Γi = 210

[Tc,iM

3i

P 4ci

](1/6)

, (C.52)

where T is in K, p is in bar, M is in g/mol and λ is in W/mK. This model yields an errorof 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-Warmeatlas [157] recommended the following method forcalculation of the mixture surface tension

σm = p2/3c,mT 1/3

c,m

(1− tr,m

am

)nm

bm , (C.53)

where

bi = 0.1196.

[1 +

Ts,ri ln(pc,m/1.01325)

1− Ts,ri

], bm =

∑xibi , (C.54)

am = 1, nm = 11/9, Tc,m =∑

j=1

xiTc,j, υc,m =∑

j

xjυc,j, Zc,j =pc,jυc,j

RTc,j

, (C.55)

Zm =∑

j

xjZc,j, pc,m =RTc,mZc,m

υc,m

, Ts,ri =Tb,i

Tc,i

, (C.56)

where Tb,i=T (p=1.01325 bar) is the normal boiling point temperature of the pure com-ponent i. T is in K, p is in bar and σ is in N/m. The Lucas and Luckas correlation yieldsan error of <5%.

C.3 Software packagesThere exists a number of software packages for the prediction of thermodynamic andtransport properties. These include:

1. ASPEN Plus (

2. CHEMCAD

3. SUPERPRO

4. REFPROP


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