course_Heat exchanger_04

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]

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Table of contents

Table of contents 1 Introduction 8 1.1 Programm outline . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8 1.2 Instructor . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 10 2 Classification of heat exchangers 2.1 Classification by construction . . . . . . . . . . . . . . . . . . . . 2.1.1 Tubular heat exchanger . . . . . . . . . . . . . . . . . . . 2.2 Double pipe heat exchanger . . . . . . . . . . . . . . . . . . . . . 2.3 Spiral tube heat exchanger . . . . . . . . . . . . . . . . . . . . . . 2.4 Shell and tube heat exchanger . . . . . . . . . . . . . . . . . . . . 2.4.1 Fixed tubesheet . . . . . . . . . . . . . . . . . . . . . . . . 2.4.2 U-tube . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.4.3 Floating head . . . . . . . . . . . . . . . . . . . . . . . . . 2.5 Plate heat exchangers . . . . . . . . . . . . . . . . . . . . . . . . 2.5.1 Gasketed plate heat exchanger . . . . . . . . . . . . . . . 2.5.2 Welded- and Brazed-Plate exchanger (W. PHE and BHE) 2.5.3 Spiral Plate Exchanger (SPHE) . . . . . . . . . . . . . . . 2.6 Extended surface . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.6.1 Plate fin . . . . . . . . . . . . . . . . . . . . . . . . . . . . 2.6.2 Tube fin . . . . . . . . . . . . . . . . . . . . . . . . . . . . 3 Code and standards 3.1 TEMA Designations . . . . . . . . . 3.2 Classification by construction STHE 3.2.1 Fixed tube sheet . . . . . . . 3.2.2 U-Tube Heat Exchanger . . . 3.2.3 Floating Head Designs . . . . 3.3 Shell Constructions . . . . . . . . . . 3.4 Tube side construction . . . . . . . . 3.4.1 Tube-Side Header: . . . . . . 3.4.2 Tube-Side Passes . . . . . . . 3.4.3 Tubes Type . . . . . . . . . . 3.4.4 Tube arrangement . . . . . . 3.4.5 Tube side passes . . . . . . . 3.5 Shell side construction . . . . . . . . 3.5.1 Shell Sizes . . . . . . . . . . . 3.5.2 Shell-Side Arrangements . . . 3.6 Baffles and tube bundles . . . . . . . 3.6.1 The tube bundle . . . . . . .

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Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

Table of contents

3.6.2 3.6.3 3.6.4 3.6.5 3.6.6

Baffle . . . . . . . . . . Vapor Distribution . . . Tube-Bundle Bypassing Tie Rods and Spacers . . Tubesheets . . . . . . .

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4 Basic Design Equations of Heat Exchangers 4.1 LMTD-Method . . . . . . . . . . . . . . . . . . 4.1.1 Logarithmic mean temperature different 4.1.2 Correction Factor . . . . . . . . . . . . . 4.1.3 Overall heat transfer coefficient . . . . . 4.1.4 Heat transfer coefficient . . . . . . . . . 4.1.5 Fouling factor (hid , hod ) . . . . . . . . . . 4.2 ε- NTU . . . . . . . . . . . . . . . . . . . . . . 4.3 Link between LMTD and NTU . . . . . . . . . 4.4 The Theta Method . . . . . . . . . . . . . . . .

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5 Thermal Design 5.1 Design Consideration . . . . . . . . . . . . . . . . 5.1.1 Fluid Stream Allocations . . . . . . . . . . 5.1.2 Shell and tube velocity . . . . . . . . . . . 5.1.3 Stream temperature . . . . . . . . . . . . 5.1.4 Pressure drop . . . . . . . . . . . . . . . . 5.1.5 Fluid physical properties . . . . . . . . . . 5.2 Design data . . . . . . . . . . . . . . . . . . . . . 5.3 Tubeside design . . . . . . . . . . . . . . . . . . . 5.3.1 Heat-transfer coefficient . . . . . . . . . . 5.3.2 Pressure drop . . . . . . . . . . . . . . . . 5.4 Shell side design . . . . . . . . . . . . . . . . . . . 5.4.1 Shell configuration . . . . . . . . . . . . . 5.4.2 Tube layout patterns . . . . . . . . . . . . 5.4.3 Tube pitch . . . . . . . . . . . . . . . . . . 5.4.4 Baffling . . . . . . . . . . . . . . . . . . . 5.4.5 Equalize cross-flow and window velocities . 5.4.6 Shellside stream analysis (Flow pattern) . 5.4.7 Heat transfer coefficient and pressure drop 5.4.8 Heat transfer coefficient . . . . . . . . . . 5.4.9 Pressure drop . . . . . . . . . . . . . . . . 5.5 Design Algorithm . . . . . . . . . . . . . . . . . . 6 Specification sheet

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Table of contents

6.1 6.2 6.3 6.4

Information included . . . . Information not included . . Operation conditions . . . . Bid evaluation . . . . . . . . 6.4.1 Factor to be consider

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7 Storage, Installation, Operation and Maintenance 7.1 Storage . . . . . . . . . . . . . . . . . . . . . . . . . 7.2 Installation . . . . . . . . . . . . . . . . . . . . . . 7.2.1 Installation Planning . . . . . . . . . . . . . 7.2.2 Installation at Jobsite . . . . . . . . . . . . 7.3 Operation . . . . . . . . . . . . . . . . . . . . . . .

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8 Heat exchanger tube side mainenance (Repair vs replacement 8.1 Introduction . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.2 Repair vs. Replace - Factors To Consider . . . . . . . . . . . . . . 8.3 Heat Exchanger maintenance Options . . . . . . . . . . . . . . . . 8.4 Repair option . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.4.1 Plug . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.4.2 Sleeving . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.4.3 Tube Expansion . . . . . . . . . . . . . . . . . . . . . . . . 8.5 Replacement option . . . . . . . . . . . . . . . . . . . . . . . . . . 8.5.1 Retubing . . . . . . . . . . . . . . . . . . . . . . . . . . . . 8.5.2 Rebundling . . . . . . . . . . . . . . . . . . . . . . . . . . 8.5.3 Complete replacement (New unit) . . . . . . . . . . . . . . 8.6 Conclusions . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 9 Troubleshooting 9.1 Heat exchangers’ problems . . . . . . . . . . . 9.2 Fouling . . . . . . . . . . . . . . . . . . . . . . 9.2.1 Costs of fouling . . . . . . . . . . . . . 9.2.2 Facts about fouling . . . . . . . . . . . 9.2.3 Types of Fouling . . . . . . . . . . . . 9.2.4 Fouling Mechanisms . . . . . . . . . . 9.2.5 Conditions Influencing Fouling . . . . . 9.2.6 Fouling control . . . . . . . . . . . . . 9.2.7 Fouling cleaning methods . . . . . . . 9.3 Leakage/Rupture of the Heat Transfer Surface 9.3.1 Cost of leakage . . . . . . . . . . . . . 9.3.2 Cause of differential thermal expansion 9.4 Corrosion . . . . . . . . . . . . . . . . . . . .

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Table of contents

9.5 9.6

9.7 9.8

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9.4.1 Corrosion effects . . . . . . . . . . . . . . . . . . . 9.4.2 Causes of corrosion . . . . . . . . . . . . . . . . . . 9.4.3 Type of corrosion . . . . . . . . . . . . . . . . . . . 9.4.4 Stress corrosion . . . . . . . . . . . . . . . . . . . . 9.4.5 Galvanic corrosion . . . . . . . . . . . . . . . . . . 9.4.6 Pitting . . . . . . . . . . . . . . . . . . . . . . . . . 9.4.7 Uniform or rust corrosion . . . . . . . . . . . . . . 9.4.8 Crevice corrosion . . . . . . . . . . . . . . . . . . . 9.4.9 Materials of Construction . . . . . . . . . . . . . . 9.4.10 Fabrication . . . . . . . . . . . . . . . . . . . . . . Troubleshooting . . . . . . . . . . . . . . . . . . . . . . . . Past failure incidents . . . . . . . . . . . . . . . . . . . . . 9.6.1 Ethylene Oxide Redistillation Column Explosion: . 9.6.2 Brittle Fracture of a Heat Exchanger . . . . . . . . 9.6.3 Cold Box Explosion . . . . . . . . . . . . . . . . . . Failure scenarios and design solutions . . . . . . . . . . . . Discussion . . . . . . . . . . . . . . . . . . . . . . . . . . . 9.8.1 Use of Potential Design Solutions Table . . . . . . . 9.8.2 Special Considerations . . . . . . . . . . . . . . . . Troubleshooting Examples . . . . . . . . . . . . . . . . . . 9.9.1 Shell side temperature uncontrolled . . . . . . . . . 9.9.2 Shell assumed banana-shape . . . . . . . . . . . . . 9.9.3 Steam condenser performing below design capacity 9.9.4 Steam heat exchanger flooded . . . . . . . . . . . .

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10 Unresolved problems in the heat exchangers design 120 10.1 Future trend . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 120 Bibliography A Heat transfer coefficient A.1 Single phase . . . . . . . . . . . . . . . . . . . . . . . . . . . . A.1.1 Inside tube: Turbulent flow . . . . . . . . . . . . . . . A.1.2 Inside tube: Laminar flow . . . . . . . . . . . . . . . . A.1.3 Shell side . . . . . . . . . . . . . . . . . . . . . . . . . A.1.4 Plate heat exchanger . . . . . . . . . . . . . . . . . . . A.2 Condensation . . . . . . . . . . . . . . . . . . . . . . . . . . . A.2.1 Condensation on vertical plate or outside vertical tube A.2.2 Condensation on external horizontal tube . . . . . . . A.2.3 Condensation on banks of horizontal tube . . . . . . . A.2.4 Condensation inside horizontal tube . . . . . . . . . . .

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A.3 Two phase flow: Pure fluid . . . . . . . . . . . . A.3.1 Steiner [140] correlation . . . . . . . . . A.3.2 Kattan et al. [77] correlation . . . . . . . A.3.3 Kandlikar [70] correlation . . . . . . . . A.3.4 Chen [19] correlation . . . . . . . . . . . A.3.5 Gungor and Winterton [52] correlation . A.3.6 Shah [130] correlation . . . . . . . . . . . A.3.7 Schrock and Grossman [129] correlation . A.3.8 Dembi et al. [30] correlation . . . . . . . A.3.9 Klimenko [84] correlation . . . . . . . . . A.3.10 Jung et al. [64] correlation . . . . . . . . A.4 Two phase flow: Mixture . . . . . . . . . . . . . A.4.1 Steiner [140] correlation . . . . . . . . . A.4.2 Kandlikar [71] correlation . . . . . . . . A.4.3 Bennett and Chen [8] correlation . . . . A.4.4 Palen [111] correlation . . . . . . . . . . A.4.5 Jung et al. [64] correlation . . . . . . . . B Pressure drop B.1 Single phase . . . . . . . . . . . . . . . . . B.2 Two phase . . . . . . . . . . . . . . . . . . B.2.1 Friedel [42] model . . . . . . . . . B.2.2 Lockhart and Martinelli [91] model B.2.3 Chisholm [22] model . . . . . . . .

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C Physical properties C.1 Physical properties: Pure fluid . . . . . . . . . . . C.1.1 Specific heat . . . . . . . . . . . . . . . . . C.1.2 Vapor pressure . . . . . . . . . . . . . . . C.1.3 Liquid viscosity . . . . . . . . . . . . . . . C.1.4 Vapor dynamic viscosity VDI-Wärmeatlas C.1.5 Dynamic viscosity of Fenghour et al. [40] . C.1.6 Surface tension . . . . . . . . . . . . . . . C.1.7 Thermal conductivity for liquids . . . . . . C.1.8 Thermal conductivity for gases . . . . . . C.1.9 Specific enthalpy . . . . . . . . . . . . . . C.2 Physical properties: Mixture . . . . . . . . . . . . C.2.1 Liquid dynamic viscosity of mixtures . . . C.2.2 Vapor dynamic viscosity of mixtures . . . C.2.3 Liquid thermal conductivity of mixtures .

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Table of contents

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


8

1 Introduction

1

Introduction

Heat exchanger is an important and expensive item of equipment that is used almost in every industry (oil and petrochemical, sugar, food, pharmaceutical and power industry). A better understanding of the basic principles of heat transfer and fluid flow and their application to the design and operation of heat exchangers that you gain from this course will enable you to improve their efficiency and extend their life. You understand how to use the applicable API, TEMA and ASME recommended practices, standards and codes for heat exchangers. This will enable you to communicate with the designers, manufacturers and bidders of heat exchangers. You will understand how to avoid fouling, corrosion and failure and leak problems by your design. You will also be able to survey and troubleshoot heat exchangers and assist in performing inspection, cleaning, and maintenance. You will be exposed to recent development and future trend in heat exchangers. The course includes worked examples to reinforce the key learning as well as a demonstration of mechanical design and challenging problems encountered in the operation of heat exchangers. Objectives • To learn the classification, code and standards (API, TEMA,...) and selection procedure for heat exchangers. • To review the thermal and mechanical design of heat exchangers. • To learn the installation, operation and maintenance procedure for heat exchanger. • To acquire information that will enable decisions to be made on the repair and refurbishment of aging equipment as well as repair vs. replacement options. • To learn techniques of failure elimination and appropriate maintenance and troubleshooting procedures. • To delineate the factors that lead to overall economically advantageous decisions. Who should attend: Project engineers, process engineers and plant engineers in the oil, chemical, sugar, power, and other industries who requires a wider and deeper appreciation of heat exchangers design, performance and operation. The detailed review of thermal and mechanical design is particularly useful to plant and maintenance engineers as well as to those generally knowledgeable in the subject, but who require a refresher or update. Codes and standards are useful for project engineer to help him communicate with manufacturers, designers and bidders of heat exchangers. Troubleshooting procedures are important for process engineers. Participants will be taken through an intensive primer of heat transfer principles as applicable to heat exchangers.

1.1

Programm outline

1. DAY I: HEAT EXCHANGERS CLASSIFICATION APPLICATION, CODE AND STANDARDS • • • • • •

Classification according to construction (tubular, plate, finned, enhanced) Classification according to service (cooler, heater, condenser, reboiler, etc..) Construction, applications, range and limitations and sizes Code and standards (TEMA, API,...) TEMA nomenclature: rear end head types, shell types, font end types TEMA standards: shell size, tube size, baffle, selection of materials, component design, nozzle loadings, supports, lifting features, high pressure, low temperature, specials designs


1.1 Programm outline

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2. DAY II HEAT TRANSFER FUNDAMENTALS AND THERMAL DESIGN • Heat transfer mechanisms: conduction and convection as related to heat exchangers • Temperature difference in heat exchanger: – LMTD Method – ε-NTU Method – θ-Method • Overall heat transfer coefficient • Heat transfer coefficient and pressure drop for single phase and multiphase (evaporation and condensation) • Resistances to fouling • Illustration examples using the software CHEMCAD 3. DAY III MECHANICAL DESIGN OF HE • Mechanical design: shells, channels and heads, tubesheets, bundles, tubestubesheet attachment • Design strategy, design algorithm • Heat exchanger: – Selection procedure – Specification sheet – Bid evaluation • Worked example (USING CHEMCAD) 4. DAY IV Storage, Installation, Operation, Maintenance • • • • • • • •

Storage Installation procedure Operation start up shut down Maintenance Cleaning Repair – Plug – Sleeving – Expansion • Replacement – Retubing – Rebundling – Replacement (new unit)

5. DAY V Troubleshooting • Heat exchangers’ problem Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

10

1 Introduction

– Fouling: causes, mechanisms, design considerations and exchanger selection, remedies, cleaning – Leakage: Location (tube sheet, tube failure), causes (differential thermal expansion, flow-induced vibration), – Corrosion: Type, causes, material of construction, fabrication – Vibration: causes (velocity), design procedure to avoid vibration including baffle selection, rod baffles, impingement baffles • Past incidents failure. • Examples of common problems encountered in heat exchangers (low rate, uncontrolled outlet temperature, failure of tubes near the inlet nozzles) Achieve the learning outcomes to: Understand the principles of heat transfer and fluid flow, application of industry practices and a substantial amount of supporting data needed for design, performance and operation of modern heat exchangers. Gain insight not only into shell and tube heat exchangers but also heat transfer fundamentals as applied to heat exchangers, the types of heat exchangers and their application, and recent advance in heat exchanger technologies Become familiar with the practical aspects and receive tips on shell and tube heat exchanger thermal design and rating: mechanical design and rating using the applicable API, TEMA and ASME recommended practices, standards and codes, troubleshooting, and performance improvement and enhancement Avoid future problems by gaining insight into vibration forcing mechanisms Enhance your awareness of causes of failure and learn practical ways for determining and correcting them Daily Schedule: 8:00 Registration and Coffee (1st day only) 8:30 Session begins 4:30 Adjournment There will be a forty-minute lunch break each day in addition to refreshment and networking break of 20 minutes during each morning and afternoon session.

1.2

Instructor

Faculty: Ali. Rabah, BSc. MSc., PhD., MSES., Assistant professor, Department of Chemical Engineering University of Khartoum Dr. Rabah holds a BSc. degree (Chemical Engineering) from the University of Khartoum, MSc. degree from university of Nairobi, Kenya, and PhD. degree from University of Hannover, Germany. He has a wide professional experience in teaching heat and mass transfer and engineering thermodynamics to BSc and MSc Chemical, Mechanical and Petroleum Engineering students. Dr. Rabah is a consultant engineer to a number of chemical industries and factories. He has developed and delivered numerous designs of heat exchangers, evaporators and boilers. He designed, for example, a 5 ton/hr (10 bar) fired tube boiler. His design is under fabrication. Dr. Rabah has designed and manufactured double pipe heat exchangers for education proposes to a number of chemical engineering departments country-wide e.g. University of Nileen. Dr. Rabah assumed engineering design positions with responsibilities covering design, construction and inspection of heat transfer equipments. The design projects are sponsored by the federal ministry of research and technology and the University of Khartoum consultancy cooperation. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

1.2 Instructor

11

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


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2

2 Classification of heat exchangers

Classification of heat exchangers

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

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ii. Drum-type (b) Fixed-matrix • Flow arrangement 1. Single pass (a) Parallel flow (b) Counter flow (c) Cross flow 2. Multipass (a) Extended surface H.E. i. Cross counter flow ii. Cross parallel flow (b) Shell and tube H.E. i. Parallel counter flow (Shell and fluid mixed, M shell pass, N Tube pass) ii. Split flow iii. Divided flow (c) Plate H.E. (N-parallel plate multipass) • Number of fluids 1. Two-fluid 2. Three fluid 3. N-fluid (N > 3) • Transfer mechanisms 1. 2. 3. 4.

Single phase convection on both sides Single phase convection on one side, two-phase convection on the other side Two-phase convection on both sides Combined convection and radiative heat transfer

• Classification based on service: Basically, a service may be single phase (such as the cooling or heating of a liquid or gas) or two-phase (such as condensing or vaporizing). Since there are two sides to an STHE, this can lead to several combinations of services. Broadly, services can be classified as follows: single-phase (both shellside and tubeside); condensing (one side condensing and the other single-phase); vaporizing (one side vaporizing and the other side single-phase); and condensing/vaporizing (one side condensing and the other side vaporizing). The following nomenclature is usually used: – Heat exchanger: both sides singlephase and process streams (that is, not a utility). – Cooler: one stream a process fluid and the other cooling water or air. Dirty water can be used as the cooling medium. The top of the cooler is open to the atmosphere for access to tubes. These can be cleaned without shutting down the cooler by removing the distributors one at a time and scrubbing the tubes. – Heater: one stream a process fluid and the other a hot utility, such as steam or hot oil. – Condenser: one stream a condensing vapor and the other cooling water or air. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

2.1

Classification by construction

The principal types of heat exchanger are listed again as 1. Tubular exchanger 2. Plate exchanger 3. Extended surface 4. Regenerative 2.1.1 Tubular heat exchanger Tubular heat exchanger are generally built of circular tubes. Tubular heat exchanger is further classified into: • Double pipe heat exchanger • Spiral tube heat exchanger • Shell and tube heat exchanger Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

2.2 Double pipe heat exchanger

2.2

15

Double pipe heat exchanger

This is usually consists of concentric pipes. One fluid flow in the inner pipe and the other fluid flow in the annulus between pipes. The two fluid may flow concurrent (parallel) or in counter current flow configuration; hence the heat exchanger are classified as: • counter current double pipe heat exchanger (see Fig. 4.1and Fig. 2.2)and • cocurrent double pipe heat exchanger

Figure 2.1. Double pipe heat exchanger. Courtesy of Perry, Chemical engineering hand book Elbew 3/4"

Tee 2"x1/2"

Galv. pipe Threaded 3/4"

Galv. pipe 2"

Union 2" Part A Cu pipe 3/4"

Tee 3/4"x1/2"

Part B

Flanged Gland 2"

Specification Sheet

Valve 3/4"

Item

pump

Double Pipe Heat Exchanger Scale: None Sheet No.1 Date: 08.12.2003 Designed by: Dr.-Ing. Ali A. Rabah

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

Microvoltmeter

1

Elbew 1/2"

4

Union 1/2"

8

1

Thermocouples (NiCr-Ni)

10

Flow meter

Bypass

Bypass

Figure 2.2. Double pipe heat exchanger (Counter current)

Double pipe heat exchanger is perhaps the simplest of all heat exchanger types. The advantages of this type are: Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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i Easily by disassembly, no cleaning problem ii Suitable for high pressure fluid, (the pressure containment in the small diameter pipe or tubing is a less costly method compared to a large diameter shell.) Limitation: The double pipe heat exchanger is generally used for the application where the total heat transfer surface area required is less than or equal to 20 m2 (215 ft2 ) because it is expensive on a cost per square meter (foot) basis.

2.3

Spiral tube heat exchanger

Spiral tube heat exchanger consists of one or more spirally wound coils fitted in a shell (Fig. 2.3). Heat transfer associated with spiral tube is higher than than that for a straight tube . In addition, considerable amount of surface area can be accommodated in a given space by spiralling. Thermal expansion is no problem but cleaning is almost impossible.

Figure 2.3. Spiral tube heat exchanger. Courtesy of The German Atlas

2.4

Shell and tube heat exchanger

Shell and tube heat exchanger is built of round tubes mounted in a cylindrical shell with the tube axis parallel to that of the shell. One fluid flow inside the tube, the other flow across and along the tubes. The major components of the shell and tube heat exchanger are tube bundle, shell, front end head, rear end head, baffles and tube sheets (Fig.2.4).

Figure 2.4. Shell and tube heat exchanger


2.4 Shell and tube heat exchanger

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The shell and tube heat exchanger is further divided into three catogaries as 1. Fixed tube sheet 2. U tube 3. Floating head 2.4.1 Fixed tubesheet A fixed-tubesheet heat exchanger (Figure 2.5) has straight tubes that are secured at both ends to tubesheets welded to the shell. The construction may have removable channel covers , bonnet-type channel covers , or integral tubesheets. The principal advantage of the fixedtubesheet construction is its low cost because of its simple construction. In fact, the fixed tubesheet is the least expensive construction type, as long as no expansion joint is required.

Figure 2.5. Fixed-tubesheet heat exchanger.

Other advantages are that the tubes can be cleaned mechanically after removal of the channel cover or bonnet, and that leakage of the shellside fluid is minimized since there are no flanged joints. A disadvantage of this design is that since the bundle is fixed to the shell and cannot be removed, the outsides of the tubes cannot be cleaned mechanically. Thus, its application is limited to clean services on the shellside. However, if a satisfactory chemical cleaning program can be employed, fixed-tubesheet construction may be selected for fouling services on the shellside. In the event of a large differential temperature between the tubes and the shell, the tubesheets will be unable to absorb the differential stress, thereby making it necessary to incorporate an expansion joint. This takes away the advantage of low cost to a significant extent. 2.4.2 U-tube As the name implies, the tubes of a U-tube heat exchanger (Figure 2.6) are bent in the shape of a U. There is only one tubesheet in a Utube heat exchanger. However, the lower cost for the single tubesheet is offset by the additional costs incurred for the bending of the tubes and the somewhat larger shell diameter (due to the minimum U-bend radius), making the cost of a U-tube heat exchanger comparable to that of a fixedtubesheet exchanger. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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The advantage of a U-tube heat exchanger is that because one end is free, the bundle can expand or contract in response to stress differentials. In addition, the outsides of the tubes can be cleaned, as the tube bundle can be removed. The disadvantage of the U-tube construction is that the insides of the tubes cannot be cleaned effectively, since the U-bends would require flexible- end drill shafts for cleaning. Thus, U-tube heat exchangers should not be used for services with a dirty fluid inside tubes.

Figure 2.6. U-tube heat exchanger.

2.4.3 Floating head The floating-head heat exchanger is the most versatile type of STHE, and also the costliest. In this design, one tubesheet is fixed relative to the shell, and the other is free to ”float” within the shell. This permits free expansion of the tube bundle, as well as cleaning of both the insides and outsides of the tubes. Thus, floating-head SHTEs can be used for services where both the shellside and the tubeside fluids are dirty-making this the standard construction type used in dirty services, such as in petroleum refineries. There are various types of floating- head construction. The two most common are the pull-through with backing device and pullthrough without backing service designs. The design (Figure 2.7) with backing service is the most common configuration in the chemical process industries (CPI). The floating-head cover is secured against the floating tubesheet by bolting it to an ingenious split backing ring. This floating-head closure is located beyond the end of the shell and contained by a shell cover of a larger diameter. To dismantle the heat exchanger, the shell cover is removed first, then the split backing ring, and then the floating-head cover, after which the tube bundle can be removed from the stationary end. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

2.5 Plate heat exchangers

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Figure 2.7. Floating head with packing service.

In the design without packing service construction (Figure 2.8), the entire tube bundle, including the floating-head assembly, can be removed from the stationary end, since the shell diameter is larger than the floating-head flange. The floatinghead cover is bolted directly to the floating tubesheet so that a split backing ring is not required. The advantage of this construction is that the tube bundle may be removed from the shell without removing either the shell or the floatinghead cover, thus reducing maintenance time. This design is particularly suited to kettle reboilers having a dirty heating medium where Utubes cannot be employed. Due to the enlarged shell, this construction has the highest cost of all exchanger types.

Figure 2.8. Floating head without packing service.

2.5

Plate heat exchangers

These exchangers are generally built of thin plates. The plate are either smooth or have some form of corrugations and they are either flat or wound in exchanger. Generally Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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theses exchanger cannot accomodate high pressure/temperature differential relative the tubular exchanger. This type of exchanger is further classified as: • Gasketed plate • Fixed plate • Spiral plate 2.5.1 Gasketed plate heat exchanger Gasketed plate heat exchanger (see Fig. 2.9) consists of a series of corrugated alloy material channel plates, bounded by elastomeric gaskets are hung off and guided by longitudinal carrying bars, then compressed by large-diameter tightening bolts between two pressure retaining frame plates (cover plates).

Figure 2.9. Plate heat exchanger

The frame and channel plates have portholes which allow the process fluids to enter alternating flow passages (the space between two adjacent-channel plates) Fig.2.10. Gaskets around the periphery of the channel plate prevent leakage to the atmosphere and also prevent process fluids from coming in contact with the frame plates. No inter fluid leakage is possible in the port area due to a dual-gasket seal. Fig.2.11 shows the plate profiles. Expansion of the initial unit is easily performed in the field without special considerations. The original frame length typically has an additional capacity of 15-20 percent more channel plates (i.e. surface area). In fact, if a known future capacity is available during fabrication stages, a longer carrying bar could be installed, and later, increasing the surface area would be easily handled. When the expansion is needed, simply untighten the carrying bolts, pull back the frame plate, add the additional channel plates, and tighten the frame plate. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


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Figure 2.10. Plate heat exchanger flow configuration

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

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Figure 2.11. Plate and frame of a plate heat exchanger

• Very close temperature approach can be obtained • low hold up volume • LMTD is fully utilized • More economical when material cost are high Disadvantages: • Low pressure <30 bar (plate deformation) • Working temperature of < (500 F) [250 o C] (maximum gasket temperature) see table 2.1.

Table 2.1. Plate Heat Exchanger Gasket Materials

Material Styrene-Butadiene Neoprene Acrylonitrile- Butadiene Ethylene/Propylene Fluorocarbon Resin-Cured Butyl Compressed Asbestos

Common name Buna-S Neoprene Buna-N EPDM Viton Resin-Cured Butyl Compressed Asbestos

Temperature limit (F) 185 250 275 300 300 300 500

2.5.2 Welded- and Brazed-Plate exchanger (W. PHE and BHE) To overcome the gasket limitations, PHE manufacturers have developed welded-plate exchangers. There are numerous approaches to this solution: weld plate pairs together with the other fluid-side conventionally gasketed, weld up both sides but use a horizonal Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


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stacking of plates method of assembly, entirely braze the plates together with copper or nickel brazing, diffusion bond then pressure form plates and bond etched, passage plates Fig. 2.12 and Fig. 2.13. Typical applications include district heating where the low cost and minimal maintenance have made this type of heat exchanger especially attractive.

Figure 2.12. Welded or blazed plate heat exchanger

Figure 2.13. Fin-Plate heat exchanger

Most methods of welded-plate manufacturing do not allow for inspection of the heattransfer surface, mechanical cleaning of that surface, and have limited ability to repair or plug off damage channels. Consider these limitations when the fluid is heavily fouling, has solids, or in general the repair or plugging ability for severe services. 2.5.3 Spiral Plate Exchanger (SPHE) The spiral-plate heat exchanger (SHE) may be one exchanger selected primarily on its virtues and not on its initial cost. SPHEs offer high reliability and on-line performance in many severely fouling services such as slurries. The SHE is formed by rolling two strips Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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of plate, with welded-on spacer studs, upon each other into clock-spring shape Fig.2.14 and Fig.2.15. This forms two passages. Passages are sealed off on one end of the SHE by welding a bar to the plates; hot and cold fluid passages are sealed off on opposite ends of the SHE. A single rectangular flow passage is now formed for each fluid, producing very high shear rates compared to tubular designs. Removable covers are provided on each end to access and clean the entire heat transfer surface.

Figure 2.14. Spiral Plate heat exchanger

Pure countercurrent flow is achieved and LMTD correction factor is essentially = 1.0. Since there are no dead spaces in a SHE, the helical flow pattern combines to entrain any solids and create high turbulence creating a self-cleaning flow passage. There are no thermal-expansion problems in spirals. Since the center of the unit is not fixed, it can torque to relieve stress. The SHE can be expensive when only one fluid requires a Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


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high alloy material. Since the heat-transfer plate contacts both fluids, it is required to be fabricated out of the higher alloy. SHEs can be fabricated out of any material that can be cold-worked and welded. The channel spacings can be different on each side to match the flow rates and pressure drops of the process design. The spacer studs are also adjusted in their pitch to match the fluid characteristics. As the coiled plate spirals outward, the plate thickness increases from a minimum of 2 mm to a maximum (as required by pressure) up to 10 mm. This means relatively thick material separates the two fluids compared to tubing of conventional exchangers.

a) Spiral flow in both channels

b) Flow are both spiral and axial

Figure 2.15. Spiral Plate heat exchanger

Applications: The most common applications that fit SHE are slurries. The rectangular channel provides high shear and turbulence to sweep the surface clear of blockage and causes no distribution problems associated with other exchanger types. A localized restriction causes an increase in local velocity which aids in keeping the unit free flowing. Only fibers that are long and stringy cause SHE to have a blockage it cannot clear itself. As an additional antifoulant measure, SHEs have been coated with a phenolic lining. This provides some degree of corrosion protection as well, but this is not guaranteed due to pinholes in the lining process. There are three types of SHE to fit different applications: • Type I is the spiral-spiral flow pattern (Fig. 2.15a). It is used for all heating and cooling services and can accommodate temperature crosses such as lean/rich services in one unit. The removable covers on each end allow access to one side at a time to perform maintenance on that fluid side. Never remove a cover with one side under pressure as the unit will telescope out like a collapsible cup. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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• Type II units are the condenser and reboiler designs (Fig. 2.15b). One side is spiral flow and the other side is in cross flow. These SHEs provide very stable designs for vacuum condensing and reboiling services. A SHE can be fitted with special mounting connections for reflux-type ventcondenser applications. The vertically mounted SHE directly attaches on the column or tank. • Type III units are a combination of the Type I and Type II where part is in spiral flow and part is in cross flow. This SHE can condense and subcool in a single unit. The unique channel arrangement has been used to provide on-line cleaning, by switching fluid sides to clean the fouling (caused by the fluid that previously flowed there) off the surface. Phosphoric acid coolers use pond water for cooling and both sides foul; water, as you expect, and phosphoric acid deposit crystals. By reversing the flow sides, the water dissolves the acid crystals and the acid clears up the organic fouling. SHEs are also used as oleum coolers, sludge coolers/ heaters, slop oil heaters, and in other services where multiple flow- passage designs have not performed well.

2.6

Extended surface

The tubular and plate exchangers described previously are all prime surface heat exchangers. The design thermal effectiveness is usually 60 % and below and the heat transfer area density is usually less than 300 m2 m3 . In many application an effectiveness of up to 90 % is essential and the box volume and mass are limited so that a much more compact surface is mandated. Usually either a gas or a liquid having a low heat transfer coefficient is the fluid on one or both sides. This results in a large heat transfer area requirements. for low density fluid (gases), pressure drop constraints tend to require a large flow area. so a question arises how can we increase both the surface area and flow area together in a reasonably shaped configuration. The surface area may be increased by the fins. The flow area is increased by the use of thin gauge material and sizing the core property. There are two most common types of extended surface heat exchangers. These are • Plate-fin • Tube-fin 2.6.1 Plate fin Plate -fin heat exchanger has fins or spacers sandwiched between parallel plates (refereed to as parting plates or parting sheets) or formed tubes as shown in fig. 2.16(left). While the plates separate the two fluid streams, the fins form the individual flow passages. Fins are used on both sides in a gas-gas heat exchanger. In gas-liquid applications fins are used in the gas side.

Figure 2.17. Finned tube heat exchanger


2.6 Extended surface

27

Figure 2.16. Examples of extended surfaces on one or both sides. Plate fins on both sides (left) and Tubes and plate fins (right).

2.6.2 Tube fin In tube fin heat exchanger, tubes of round, rectangular, or elliptical shape are generally used. Fins are generally used on the outside and also used inside the tubes in some applications. they are attached to the tube by tight mechanical fit, tension wound, gluing, soldering, brazing, welding or extrusion. Tube fin exchanger is shown in Fig. 2.16(right) and Fig.2.17


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3 Code and standards

3

Code and standards

The objective of codes and standards are best described by ASME: The objectives of code rules and standards (apart from fixing dimensional values) is to achieve minimum requirements for safe construction, in other words, to provide public protection by defining those materials, design, fabrication and inspection requirements; whose omission may radically increase operating hazards.... Experience with code rules has demonstrated that the probability of disastrous failure can be reduced to the extremely low level necessary to protect life and property by suitable minimum requirements and safety factors. Obviously, it is impossible for general rules to anticipate other than conventional service,.... Suitable precautions are therefore entirely the responsibility of the design engineer guided by the needs and specifications of the user. Over years a number of standardization bodies have been developed by individual country, manufacturers and designers to lay down nomenclatures for the size and type of shell and tube heat exchangers. These include among other • TEMA standards (Tubular Exchanger Manufacturer Association., 1998)[147] • HEI standards (Heat Exchanger Institute, 1980), • API (American Petroleum Institute). • Other national standards include the German (DIN), Japan, India, to mention a few. In this work, being most widely used one, the TEMA standard is presented.

3.1

TEMA Designations

In order to understand the design and operation of the shell and tube heat exchanger, it is important to know the nomenclature and terminology used to describe them and the various parts that go to their construction. Only then we can understand the design and reports given by the researchers, designers, manufacturer and users. It is essential for the designer to have a good working knowledge of the mechanical features of STHEs and how they influence thermal design. The principal components of an STHE are: • shell; • shell cover; • tubes; • channel; • channel cover; • tubesheet; • baffles; and • nozzles. Other components include tie-rods and spacers, pass partition plates, impingement plate, longitudinal baffle, sealing strips, supports, and foundation. Table 3.1 shows the nomenclature used for different parts of shell and tube exchanger in accordance with TEMA standards; the numbers refer to the feature shown in Fig. 3.2 to Fig. 3.8. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

3.1 TEMA Designations

29

Table 3.1. TEAM notations

Index Notation 1 2 3 4 5 6 7 8 9 10 11 12 13 14 15 16 17 18 19

Index

stationary head- channel 20 stationary head- bonnet 21 stationary head flange-chennel or bonnet 22 channel cover 23 stationary head - nozzle 24 stationary tube sheet 25 tubes 26 shell 27 shell cover 28 shell flange-stationary head end 29 shell flange-rear head end 30 shell nozzle 31 shell cover flange 32 expansion joint 33 floating tube sheet 34 floating head cover 35 floating head flange 36 floating head backing device 37 split shear ring 38 39

Notation slip on backing flange floating head cover-external floating tube sheet skirt packing box packing packing gland latern ring tie rods and spacers traverse baffle or support plate impingement plate longitudinal baffle pass partition vent connection drain connection instrument connection support saddle lifting lug support bracket weir liquid level connection

Because of the number of variations in mechanical designs for front and rear heads and shells, and for commercial reasons, TEMA has divided STHE into main three components: front head, shell and rear head. Fig. 3.1 illustrates TEMA nomenclature for the various construction possibilities. TEMA has classified the front head channel and bonnet types as given the letters (A,B,C,N,D) and the shell is classified according to the nozzles locations for the inlet and outlet. There are type of shell configuration ( E,F,G,H,J,K,X). Similarly the rear head is classified ( M,N,P,S,T,U,W). Exchangers are described by the letter codes of the three sections. The first letter stands for the front head, the second letter for the shell type and the third letter for the rear head type. For example a BFL exchanger has a bonnet cover, two-shell pass with longitudinal baffles and a fixed tube sheet rear head. In addition to these the size of the exchanger is required to be identified with the notation. The size is identified by the shell inside diameter (nominal) and tube length (both are rounded to the nearest integer in inch or mm). Demonstration examples are shown below: • Type AES size 23-192 in (590-4880): This exchanger has a removable channel cover (A), single pass shell (E) and Split ring floating front head (S) it has , 23 in (590 mm) inside diameter with tubes of 16 ft (4880 mm) long. • Type BGU Size 19-84 (480-2130)This exchanger has a bonnet-type stationary front head (B), split flow shell (G) and U-tube bundle rear head(U) with 19 in (480) inside diameter and 7 ft (2130 mm) tube length. • Type AFM size 33-96 (840-2440): This exchanger has a removable channel and cover front head (A), two-pass shell (F) and fixed tube sheet bonnet-type rear head (M) with 331/8 in (840 mm) inside diameter and 8ft (2440 mm) tube length. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

30


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

In the above illustration the term single pass and two pass shell have been used. This mean that the shell side fluid travels only one through the shell (single pass) or twice (two pass shell). Two pass shell mean that the fluid enters at one end, travel to other end and back to the end where it entered (making U-turn). Similarly there are multiple pases. To be remembered is that the number of tube passes is equal to or greater than the number of shell passes. Generally the multi shell and tube passes are usually designated by two numerals separated by a hyphen, with the first numeral indication the number of shell pass and the other stands for the tube passes. For example a one-shell pass and two tube pass AEL exchanger will be written as 1-2 AEL. To be remembered is that this not an TEMA standards. TEMA requires the number of shell and tube passes to be spelled out Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

3.1 TEMA Designations

31

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


3 Code and standards 32

Type of design T.E.M.A. rear -head type Relative cost increases from A (least expensive) through E (most expensive) Provision for differential expansion

Removable bundle Replacement bundle possible Individual tubes replaceable Tube cleaning by chemicals inside and outside Interior tube cleaning mechanically Exterior tube cleaning mechanically: Triangular pitch No Square pitch Hydraulic-jet cleaning: Tube interior Tube exterior Double tube sheet feasible Number of tube passes Internal gaskets eliminated

Yes

Yes

Yes

No

Expansion joint in shell No

B

L or M or N

Fixed tube sheet

No Yes

Special tools required

Yes

Only those in outside row

Yes

Yes

Individual tubes free to expand

A

U

U-tube

No Yes

Yes

Yes

Yes

Yes

Yes

Floating head

C

W

Packed lantern-ring floating head

No Yes

Yes

Yes

Yes

Yes

Yes

Floating head

E

S

Internal floating head (split backing ring)

No Yes

Yes

Yes

Yes

Yes

Yes

Floating head

D

P

Outside packed floating head

Yes

Yes

Yes

Yes

Yes

Yes

Floating head

E

T

Pull-through floating head

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

No No

No No practical limitations

Yes Yes

Yes No practical limitations

No

Yes Yes No No practical limitations

Yes

Yes Yes No Limited to one or two passes

No

Yes Yes Yes Any even number possible

Yes

Special tools required Yes Yes No practical limitations

Yes

Yes No

Yes


3.2 Classification by construction STHE

3.2

33

Classification by construction STHE

Fig. 3.2 to Fig. 3.8 show details of the construction of the TEMA types of shell-and-tube heat exchangers. These types are: • Fixed tube sheet • U-tube • Floating head 3.2.1 Fixed tube sheet Fixed-tube-sheet exchangers (Fig. 3.2) are used more often than any other type, and the frequency of use has been increasing in recent years. The tube sheets are welded to the shell. Usually these extend beyond the shell and serve as flanges to which the tube-side headers are bolted. This construction requires that the shell and tube-sheet materials be weldable to each other. When such welding is not possible, a blind-gasket type of construction is utilized. The blind gasket is not accessible for maintenance or replacement once the unit has been constructed. This construction is used for steam surface condensers, which operate under vacuum.

Figure 3.2. Heat-exchanger-component nomenclature. Fixed tube heat sheet shell and tube heat exchanger. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)

The tube-side header (or channel) may be welded to the tube sheet, as shown in Fig. 3.1 for type C and N heads. This type of construction is less costly than types B and M or A and L and still offers the advantage that tubes may be examined and replaced without disturbing the tube-side piping connections. There is no limitation on the number of tube-side passes. Shell-side passes can be one or more, although shells with more than two shell side passes are rarely used. Tubes can completely fill the heat-exchanger shell. Clearance between the outermost tubes and the shell is only the minimum necessary for fabrication. Between the inside of the shell and the baffles some clearance must be provided so that baffles can slide into the shell. Fabrication tolerances then require some additional clearance between the outside of the baffles and the outermost tubes. The edge distance between the outer tube limit (OTL) and the baffle diameter must be sufficient to prevent vibration of the tubes from breaking through the baffle holes. The outermost tube must be contained within the OTL. Clearances between the inside shell diameter and OTL are 13 mm (1/2 in) for 635-mm(25-in-) inside-diameter shells and up, 11 mm for 254- through 610-mm (10- through 24-in) pipe shells, and slightly less for smaller-diameter pipe shells. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

34


Tubes can be replaced. Tube-side headers, channel covers, gaskets, etc., are accessible for maintenance and replacement. Neither the shell-side baffle structure nor the blind gasket is accessible. During tube removal, a tube may break within the shell. When this occurs, it is most difficult to remove or to replace the tube. The usual procedure is to plug the appropriate holes in the tube sheets. Differential expansion between the shell and the tubes can develop because of differences in length caused by thermal expansion. Various types of expansion joints are used to eliminate excessive stresses caused by expansion. The need for an expansion joint is a function of both the amount of differential expansion and the cycling conditions to be expected during operation. A number of types of expansion joints are available (Fig. 3.3)

Figure 3.3. Expansion joints.

. a Flat plates. Two concentric flat plates with a bar at the outer edges. The flat plates can flex to make some allowance for differential expansion. This design is generally used for vacuum service and gauge pressures below 103 kPa (15 lbf/in2). All welds are subject to severe stress during differential expansion. b Flanged-only heads. The flat plates are flanged (or curved). The diameter of these heads is generally 203 mm (8 in) or more greater than the shell diameter. The welded joint at the shell is subject to the stress referred to before, but the joint Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


35

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

36


that for a fixed-tube-sheet exchanger because of limitations on bending tubes of a very short radius.

Figure 3.4. Heat-exchanger-component nomenclature. U-tube heat exchanger. Type CFU. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)

The U-tube design offers the advantage of reducing the number of joints. In high-pressure construction this feature becomes of considerable importance in reducing both initial and maintenance costs. The use of U-tube construction has increased significantly with the development of hydraulic tube cleaners, which can remove fouling residues from both the straight and the U-bend portions of the tubes. Rods and conventional mechanical tube cleaners cannot pass from one end of the U tube to the other. Power-driven tube cleaners, which can clean both the straight legs of the tubes and the bends, are available. Hydraulic jetting with water forced through spray nozzles at high pressure for cleaning tube interiors and exteriors of removal bundles is reported in the recent ASME publications. U-tube can be used for high pressure and high temperature application like kettle reboiler, evaporator, tank section heaters ,etc. The tank suction heater, as illustrated in Fig. 3.5, contains a U-tube bundle. This design is often used with outdoor storage tanks for heavy fuel oils, tar, molasses, and similar fluids whose viscosity must be lowered to permit easy pumping. Uusally the tube-side heating medium is steam. One end of the heater shell is open, and the liquid being heated passes across the outside of the tubes. Pumping costs can be reduced without heating the entire contents of the tank. Bare tube and integral low-fin tubes are provided with baffles. Longitudinal fin-tube heaters are not baffled. Fins are most often used to minimize the fouling potential in these fluids. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


37

Figure 3.5. Heat-exchanger-component nomenclature. U-tube heat exchanger. Type CFU. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)

Kettle-type reboilers, evaporators, etc. , are often U-tube exchangers with enlarged shell sections for vapor-liquid separation (Fig.3.6). The U-tube bundle replaces the floatingheat bundle of Fig. 3.4.

Figure 3.6. Kettle reboiler

The U-tube exchanger with copper tubes, cast-iron header, and other parts of carbon steel is used for water and steam services in office buildings, schools, hospitals, hotels, etc. Nonferrous tube sheets and admiralty or 90-10 copper-nickel tubes are the most frequently used substitute materials. These standard exchangers are available from a number of manufacturers at costs far below those of custombuilt process-industry equipment. 3.2.3 Floating Head Designs In an effort to reduce thermal stresses and provide a means to remove the tube bundle for cleaning, several floating rear head designs have been established. The simplest is a Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

38


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



39

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

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

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

There are no limitations upon the number of tube-side passes or upon the tube-side design pressure and temperature. The outside-packed floating-head exchanger was the most commonly used type of removable- bundle construction in chemical-plant service. The floating-tube-sheet skirt, where in contact with the rings of packing, has fine machine finish. A split shear ring is inserted into a groove in the floating-tube-sheet skirt. A slipon backing flange, which in service is held in place by the shear ring, bolts to the external floating- head cover. The floating-head cover is usually a circular disk. With an odd number of tube-side passes, an axial nozzle can be installed in such a floating- head cover. If a side nozzle is required, the circular disk is replaced by either a dished head or a channel barrel (similar to Fig. 11-36f ) bolted between floating-head cover and floating-tube-sheet skirt. The outer tube limit approaches the inside of the skirt but is farther removed from the inside of the shell than for any of the previously discussed constructions. Clearances Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

40


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

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

Figure 3.10. Heat-exchanger-component nomenclature. Exchanger with packed floating tube sheet and lantern ring. Type AES. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)

Pull-Through Floating-Head Exchanger: (Fig. 3.12) Construction is similar to that of the internal-floating-head split-backing ring exchanger except that the floating-head Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

3.3 Shell Constructions

41

cover bolts directly to the floating tube sheet. The tube bundle can be withdrawn from the shell without removing either shell cover or floating-head cover. This feature reduces maintenance time during inspection and repair. The large clearance between the tubes and the shell must provide for both the gasket and the bolting at the floating-head cover. This clearance is about 2 to 2.5 times that required by the split-ring design. Sealing strips or dummy tubes are often installed to reduce bypassing of the tube bundle.

Figure 3.11. Heat-exchanger-component nomenclature. Kettle-type floating-head reboiler. Type AKT. (Standard of Tubular Exchanger Manufacturers Association, 6th ed., 1978.)

3.3

Shell Constructions

• The most common TEMA shell type is the E shell as it is most suitable for most industrial process cooling applications. However, for certain applications, other shells offer distinct advantages. For example, the TEMA-F shell design provides for a longitudinal flow plate to be installed inside the tube bundle assembly. This plate causes the shell fluid to travel down one half of the tube bundle, then down the other half, in effect producing a counter-current flow pattern which is best for heat transfer. This type of construction can be specified where a close approach temperature is required and when the flow rate permits the use of one half of the shell at a time. In heat recovery applications, or where the application calls for increased thermal length to achieve effective overall heat transfer, shells can be installed with the flows in series. Up to six shorter shells in series is common and results in counter-current flow close to performance as if one long shell in a single pass design were used. • TEMA G and H shell designs are most suitable for phase change applications where the bypass around the longitudinal plate and counter-current flow is less important than even flow distribution. In this type of shell, the longitudinal plate offers better flow distribution in vapor streams and helps to flush out non-condensable. They are frequently specified for use in horizontal thermosiphon reboilers and total condensers. • TEMA J Shells are typically specified for phase change duties where significantly reduced shell side pressure drops are required. They are commonly used in stacked sets with the single nozzles used as the inlet and outlet. A special type of J-shell is used for flooded evaporation of shell side fluids. A separate vapor disengagement vessel without tubes is installed above the main J shell with the vapor outlet at the top of this vessel. The Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

42


• TEMA K shell, also termed a kettle reboiler, is specified when the shell side stream will undergo vaporization. The liquid level of a K shell design should just cover the tube bundle, which fills the smaller diameter end of the shell. This liquid level is controlled by the liquid flowing over a weir at the far end of the entrance nozzle. The expanded shell area serves to facilitate vapor disengagement for boiling liquid in the bottom of the shell. To insure against excessive liquid carry-though with the vapor stream, a separate vessel as described above is specified. Liquid carry-through can also be minimized by installing a mesh demister at the vapor exit nozzle. U-bundles are typically used with K shell designs. K shells are expensive for high pressure vaporization due to shell diameter and the required wall thickness. • The TEMA X shell, or crossflow shell is most commonly used in vapor condensing applications, though it can also be used effectively in low pressure gas cooling or heating. It produces a very low shell side pressure drop, and is therefore most suitable for vacuum service condensing. In order to assure adequate distribution of vapors, X-shell designs typically feature an area free of tubes along the top of the exchanger. It is also typical to design X shell condensers with a flow area at the bottom of the tube bundle to allow free condensate flow to the exit nozzle. Careful attention to the effective removal of non-condensables is vital to X-shell constructions.

3.4

Tube side construction

3.4.1 Tube-Side Header: The tube-side header (or stationary head) contains one or more flow nozzles. • The bonnet (Fig. 3.1B) bolts to the shell. It is necessary to remove the bonnet in order to examine the tube ends. The fixed-tubesheet exchanger of Fig. 3.1b has bonnets at both ends of the shell. • The channel (Fig. 3.1A) has a removable channel cover. The tube ends can be examined by removing this cover without disturbing the piping connections to the channel nozzles. The channel can bolt to the shell as shown in Fig. 3.1a and c. The Type C and Type N channels of Fig. 3.1 are welded to the tube sheet. This design is comparable in cost with the bonnet but has the advantages of permitting access to the tubes without disturbing the piping connections and of eliminating a gasketed joint. • Special High-Pressure Closures (Fig. 3.1D) The channel barrel and the tube sheet are generally forged. The removable channel cover is seated in place by hydrostatic pressure, while a shear ring subjected to shearing stress absorbs the end force. For pressures above 6205 kPa (900 lbf/in2) these designs are generally more economical than bolted constructions, which require larger flanges and bolting as pressure increases in order to contain the end force with bolts in tension. Relatively light-gauge internal pass partitions are provided to direct the flow of tube-side fluids but are designed only for the differential pressure across the tube bundle. 3.4.2 Tube-Side Passes Most exchangers have an even number of tube-side passes. The fixed-tube-sheet exchanger (which has no shell cover) usually has a return cover without any flow nozzles as shown in Fig. 3.1M; Types L and N are also used. All removable-bundle designs (except for the U tube) have a floating-head cover directing the flow of tube-side fluid at the floating tube sheet. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

3.4 Tube side construction

43

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

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Figure 3.12. Tube thickness

2. Finned tube: As the name implies, finned tube have fins to the tubular surface. Fins can be longtiudinal, radial or helical and may be on the outside or inside or on both sides of the tube. Fig. 5.7shows some of the commonly used fins. The fins are generally used when at least one of the fluid is gas. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

3.4 Tube side construction

45

Figure 3.13. Examples of extended surfaces on one or both sides. (a) Radial fins. (b) Serrated radial fins. (c) Studded surface. (d) Joint between tubesheet and low fin tube with three times bare surface. (e) External axial fins. ( f ) Internal axial fins. (9) Finned surface with internal spiral to promote turbulence. (h) Plate fins on both sides. (i) Tubes and plate fins.

(a) Integrally finned tube, which is available in a variety of alloys and sizes, is being used in shell-and-tube heat exchangers. The fins are radially extruded from thick-walled tube to a height of 1.6 mm (1/16 in) spaced at 1.33 mm (19 fins per inch) or to a height of 3.2 mm (1/8 in) spaced at 2.3 mm (11 fins per inch). External surface is approximately 2 1/2 times the outside surface of a bare tube with the same outside diameter. Also available are 0.93-mm- (0.037in-) high fins spaced 0.91 mm (28 fins per inch) with an external surface about 3.5 times the surface of the bare tube. Bare ends of nominal tube diameter are provided, while the fin height is slightly less than this diameter. The tube can be inserted into a conventional tube bundle and rolled or welded to the tube sheet by the same means, used for bare tubes. An integrally finned tube rolled Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

46


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

3.4.4 Tube arrangement The tubes in an exchanger are usually arranged in an equilateral triangular, aquare or rotated square pattern see fig.3.14. The triangular and rotated square pattern give higher heat transfer rates, but at the expenses of higher pressure drop than the the square pattern. Square or rotated square are used for hihger fouling fluid, where it is necessary to mechanically clean the outside of the tubes. The recommend tube pitch is Pt = 1.25do . Where square pattern is used for easer of cleaning, the recommended minimum clearance between the tubes is 0.25 in (6.4 mm) Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

3.5 Shell side construction

47

d

o

pt

do

pt

pt

Flow

Square pitch

Equilateral triangular pitch

Rotaed square

Figure 3.14. Tube patterns.

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

1

2 Two tube passes 1 2 3 4 Four tube passes 1

1 3

2 5

4

3

2 5

4 6

6 Six tube passes

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

3.5

Shell side construction

3.5.1 Shell Sizes Heat-exchanger shells are generally made from standard- wall steel pipe in sizes up to 305-mm (12-in) diameter; from 9.5-mm (3/8 in) wall pipe in sizes from 356 to 610 mm (14 to 24 in); and from steel plate rolled at discrete intervals in larger sizes. Clearances between the outer tube limit and the shell are discussed elsewhere in connection with the different types of construction. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

48


3.5.2 Shell-Side Arrangements 1. The one-pass shell (Fig. 3.1E) is the most commonly used arrangement. Condensers from single component vapors often have the nozzles moved to the center of the shell for vacuum and steam services. Solid longitudinal baffle is provided to form a two-pass shell (Fig. 3.1F). It may be insulated to improve thermal efficiency. (See further discussion on baffles). 2. A two-pass shell can improve thermal effectiveness at a cost lower than for two shells in series. 3. For split flow (Fig. 3.1G), the longitudinal baffle may be solid or perforated. The latter feature is used with condensing vapors. 4. double-split-flow design is shown in Fig. 3.1H. The longitudinal baffles may be solid or perforated. 5. The divided flow design (Fig. 3.1J), mechanically is like the one-pass shell except for the addition of a nozzle. Divided flow is used to meet low-pressure-drop requirements. The kettle reboiler is shown in Fig. 3.1K. When nucleate boiling is to be done on the shell-side, this common design provides adequate dome space for separation of vapor and liquid above the tube bundle and surge capacity beyond the weir near the shell cover.

3.6

Baffles and tube bundles

3.6.1 The tube bundle Tube bundle is the most important part of a tubular heat exchanger. The tubes generally constitute the most expensive component of the exchanger and are the one most likely to corrode. Tube sheets, baffles, or support plates, tie rods, and usually spacers complete the bundle. 3.6.2 Baffle Baffles are used to direct the side and tube side flows so that the fluid velocity is increased to obtain higher heat transfer rate and reduce fouling deposits. In horizontal units baffle are used to provide support against sagging and vibration damage. There are different types of baffles: 1. segemntal 2. disc and doughnut 3. orifice 4. rod type 5. nest type 6. longitudinal 7. impingment 1. Segmental Baffles Segmental or cross-flow baffles are standard. Single, double, and triple segmental baffles are used. Baffle cuts are illustrated in Fig. 3.16a. The double segmental baffle reduces crossflow velocity for a given baffle spacing. The triple segmental baffle reduces both cross-flow and long-flow velocities and has been identified as the window-cut baffle. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

3.6 Baffles and tube bundles

49

a

b

c

d

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

Minimum baffle spacing is generally one-fifth of the shell diameter and not less than 50.8 mm (2 in). Maximum baffle spacing is limited by the requirement to provide adequate support for the tubes. The maximum unsupported tube span in inches equals 74d0.75 (where d is the outside tube diameter in inches). The unsupported tube span is reduced by about 12 percent for aluminum, copper, and their alloys. Baffles are provided for heat-transfer purposes. When shell-side baffles are not required for heat-transfer purposes, as may be the case in condensers or reboilers, tube supports are installed. Maximum baffle cut is limited to about 45 percent for single segmental baffles so that every pair of baffles will support each tube. Tube bundles are generally provided with baffles cut so that at least one row of tubes passes through all the baffles or support plates. These tubes hold the entire bundle together. In pipe-shell exchangers with a horizontal baffle cut and a horizontal pass rib for directing tube Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

50


side flow in the channel, the maximum baffle cut, which permits a minimum of one row of tubes to pass through all baffles, is approximately 33 percent in small shells and 40 percent in larger pipe shells. Maximum shell-side heat-transfer rates in forced convection are apparently obtained by cross-flow of the fluid at right angles to the tubes. In order to maximize this type of flow some heat exchangers are built with segmental-cut baffles and with no tubes in the window (or the baffle cutout). Maximum baffle spacing may thus equal maximum unsupported-tube span, while conventional baffle spacing is limited to one-half of this span. The maximum baffle spacing for no tubes in the window of single segmental baffles is unlimited when intermediate supports are provided. These are cut on both sides of the baffle and therefore do not affect the flow of the shell-side fluid. Each support engages all the tubes; the supports are spaced to provide adequate support for the tubes. 2. Rod Baffles Rod or bar baffles (fig. 3.17) have either rods or bars extending through the lanes between rows of tubes. A baffle set can consist of a baffle with rods in all the vertical lanes and another baffle with rods in all the horizontal lanes between the tubes. The shell-side flow is uniform and parallel to the tubes. Stagnant areas do not exist. One device uses four baffles in a baffle set. Only half of either the vertical or the horizontal tube lanes in a baffle have rods. The new design apparently provides a maximum shell-side heat-transfer coefficient for a given pressure drop.

Figure 3.17. Rod baffles.

3. Impingement Baffle The tube bundle is customarily protected against impingement by the incoming fluid at the shell inlet nozzle when the shell-side fluid is at a high velocity, is condensing, or is a twophase fluid. Minimum entrance area about the nozzle is generally equal to the inlet nozzle area. Exit nozzles also require adequate area between the tubes and the nozzles. A full bundle without any provision for shell inlet nozzle area can increase the velocity of the inlet fluid by as much as 300 percent with a consequent loss in pressure. Impingement baffles are generally made of rectangular plate, although circular plates (Fig. 3.18) are more desirable. Rods and other devices are sometimes used to protect the tubes from impingement. In order to maintain a maximum tube count Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


51

the impingement plate is often placed in a conical nozzle opening or in a dome cap above the shell. Impingement baffles or flow-distribution devices are recommended for axial tubeside nozzles when entrance velocity is high. (a)

(c)

(B)

(d)

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

4. Longitudinal Flow Baffles In fixed-tube-sheet construction with multipass shells, the baffle is usually welded to the shell and positive assurance against bypassing results. Removable tube bundles have a sealing device between the shell and the longitudinal baffle. Flexible light-gauge sealing strips and various packing devices have been used. Removable U-tube bundles with four tube-side passes and two shell-side passes can be installed in shells with the longitudinal baffle welded in place. In split-flow shells the longitudinal baffle may be installed without a positive seal at the edges if design conditions are not seriously affected by a limited amount of bypassing. Fouling in petroleum-refinery service has necessitated rough treatment of tube bundles during cleaning operations. Many refineries avoid the use of longitudinal baffles, since the sealing devices are subject to damage during cleaning and maintenance operations. 3.6.3 Vapor Distribution Relatively large shell inlet nozzles, which may be used in condensers under low pressure or vacuum, require provision for uniform vapor distribution. 3.6.4 Tube-Bundle Bypassing Shell-side heat-transfer rates are maximized when bypassing of the tube bundle is at a minimum. The most significant bypass stream is generally between the outer tube limit and the inside of the shell. The clearance between tubes and shell is at a minimum for fixed-tube-sheet construction and is greatest for straight-tube removable bundles. Arrangements to reduce tube-bundle bypassing include: Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

52


1. Dummy tubes. These tubes do not pass through the tube sheets and can be located close to the inside of the shell. 2. Tie rods with spacers. These hold the baffles in place but can be located to prevent bypassing. 3. Sealing strips. These longitudinal strips either extend from baffle to baffle or may be inserted in slots cut into the baffles. 4. Dummy tubes or tie rods with spacers may be located within the pass partition lanes (and between the baffle cuts) in order to ensure maximum bundle penetration by the shell-side fluid. When tubes are omitted from the tube layout to provide entrance area about an impingement plate, the need for sealing strips or other devices to cause proper bundle penetration by the shell-side fluid is increased. 3.6.5 Tie Rods and Spacers Tie rods are used to hold the baffles in place with spacers, which are pieces of tubing or pipe placed on the rods to locate the baffles. Occasionally baffles are welded to the tie rods, and spacers are eliminated. Properly located tie rods and spacers serve both to hold the bundle together and to reduce bypassing of the tubes. In very large fixed-tube-sheet units, in which concentricity of shells decreases, baffles are occasionally welded to the shell to eliminate bypassing between the baffle and the shell. Metal baffles are standard. Occasionally plastic baffles are used either to reduce corrosion or in vibratory service, in which metal baffles may cut the tubes.

baffle

Rods

Spacer

Tube plate

Figure 3.19. Baffle spacers and tie rods.

3.6.6 Tubesheets Tubesheets are usually made from a round flat piece of metal with holes drilled for the tube ends in a precise location and pattern relative to one another. Tube sheet materials range as tube materials. Tubes are attached to the tube sheet by pneumatic or hydraulic pressure or by roller expansion. Tube holes can be drilled and reamed and can be machined with one or more grooves. This greatly increases the strength of the tube joint. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


53

0.4mm

3 mm

a

b

c

Figure 3.20. Tube sheet joint

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

54


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


55

4

Basic Design Equations of Heat Exchangers

There are two types of design problems: sizing and rating. In sizing the main objective is to find the geometry of the heat exchanger. Rating is to find the duty or performance for a given geometry. RATING Given: Geometry mh , Ch , Th1 , ∆ph mc , Cc , Tc1 , ∆pc Find: Q(Duty)

SIZING Given: Q(duty) mh , Ch , Th1 , ∆ph mc , Cc , Tc1 , ∆pc Find: Geometry

The are three design approaches generally used in the design of heat exchanger. These are • LMTD-method, • NTU-ε-method and • θ-method. These notation are explained in the respective sections.

4.1

LMTD-Method

Assumptions • Steady state flow (mh , mc ) • Constant overall heat transfer coefficient (U ) • Constant specific heat (Cph , Cpc ) • negligible heat loss to surrounding Heat Transfer (or rate equation) Q = U A∆Tlm F

(4.1)

where Q= heat transferred per unit time W (duty) U= overall heat transfer coefficient A= heat transfer area ∆Tlm = logarithmic mean temperature difference F = temperature correction factor Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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:

Thi

∆T1 Tho ∆T2 Tco

∆T1

Thi

∆T1

Tco

Thi Tho

Tho

Tci

Tco

∆T2

Tc

Tci

Thi

Tci Tci

Tci

Tco

Tco

Tci

Thi

Tho

Thi

Tho

Cocurrent

Tco Shell and Tube

Counter current

Tho

Figure 4.1. Temperature distribution

Cocurrent Counter current Shell and tube Plate heat exchanger

∆T1 Thi − Tci Thi − Tco Thi − Tco Thi − Tco

∆T2 Tho − Tco Tho − Tci Tho − Tci Tho − Tci

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

1. counter current

2. co-current

Solution Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

4.1 LMTD-Method

57

Thi=110 oC ∆T1=45oC Tco=65oC

Thi=110 oC Tho=75oC ∆T1 =75

∆T2

Tho=75oC

=10oC

∆T2=40oC

Tco=65oC Tci= 35 oC

Tci= 35 oC

Tci

Tco

Tco

Thi

Tho

Thi

Tci Tho Counter current

Cocurrent

Figure 4.2. Temperature distribution

1. counter current (see Fig.4.2) ∆Tlm =

∆T2 − ∆T1 10 − 75 = = 32.26o C ln(∆T2 /∆T1 ) ln(10/75)

(4.3)

∆T2 − ∆T1 40 − 45 = = 42.45o C ln(∆T2 /∆T1 ) ln(40/45)

(4.4)

2. co-current (see Fig.4.2) ∆Tlm =

4.1.2 Correction Factor • For double pipe heat exchanger

F =1

(4.5)

• Shell and tube heat exchanger. For a 1 shell 2 tube pass exchanger the correction factor is given by: q h i 1−S (R2 + 1) ln 1−RS h i  F = (4.6) √ (R − 1) ln where R=

 2−S R+1− (R2 +1)  h i √  2−S R+1− (R2 +1) 

T1 − T2 , t2 − t1

S=

t2 − t1 T1 − t 1

(4.7)

or in words R=

Range of shell f luid , Range of tube f luid

S=

Range of tube f luid M aximum temperature dif f erence

(4.8)

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

F =

i

ln 1−S √ 1−RS √ 2/S−1−R+(2/S) (1−S)(1−RS)+ R2 +1 √ √ ln 2 R2+1 2(R−1)

2/S−1−R+(2/S)

(4.9)

(1−S)(1−RS)− R +1

These equations are plotted on fig.4.4 Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

58


Example 1 For example calculate the correction factor for

1. 1-2 shell and tube heat exchanger and

2. 2-4 shell and tube heat exchanger

using the equation and the graph. T1 = 35o C, T2 = 65o C, t1 = 110o C, t2 = 75o C

R=

T1 − T2 35 − 65 = = 0.86, t2 − t1 75 − 110

S=

t2 − t1 75 − 110 = = 0.467 T1 − t1 35 − 110

(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

1-2 Shell and Tube T1 t1

2-4 Shell and Tube T1 t1

t2 T2 t2 T2 Figure 4.3. Temperature distribution for 1-2 and 2-4 shell and tube heat exchanger


4.1 LMTD-Method

59

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

4.1.3 Overall heat transfer coefficient Typical values of the overall heat transfer coefficient for various types of heat exchnager are given in . More expensive data can be found in in The determination of U is often tedious and needs data not yet available in preliminary stages of the design. Therefore, typical values of U are useful for quickly estimating the required surface area. The literature has many tabulations of such typical coefficients for commercial heat transfer services. Following is a table 4.1 with values for different applications and heat exchanger types. More values can be found in the books as [29],[127], [113], [79], [93] and [14] The ranges given in the table are an indication for the order of magnitude. Lower values are for unfavorable conditions such as lower flow velocities, higher viscosities, and additional fouling resistances. Higher values are for more favorable conditions. Coefficients of actual equipment may be smaller or larger than the values listed. Note that the values should not be used as a replacement of rigorous methods for the final design of heat exchangers, although they may serve as a useful check on the results obtained by these Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

60


methods.

Table 4.1. Typical overall coefficient

Hot Fluid Heat exchangers Water Organic solvents light oils heavy oils Gases Coolers Organic solvents light oils heavy oils gase organic solvent water Gases Heaters Steam Steam Steam Steam Steam Dowtherm Dowtherm flue gases flue gases Condensers Aqueous vapor Organic vapor Organic (some non condensable gases) Vacuum condensers Vaporizers Steam Steam Steam

Cold fluid

U (W/m2 o C)

Water organic solvent light oils heavy oils gass

800-1500 100-300 100-400 50-300 10-50

water water water60-900 water brine brine Brine

250-750 350-900 20-300 150-500 600-1200 15-250

Water organic solvent light oils heavy oils gass Heavy oils Gases steam hydrocarbon vapor

1500-4000 500-1000 300-900 60-450 30-300 50-300 20-200 30-100 30-100

water Water Water Water

1000-1500 700-1000 500-700 200-500

Aqueuos solutions Light organics Heavy organics

1000-1500 900-1200 600-900

Alternatively the overall heat transfer coefficient is evalauted from the individual heat transfer coefficient as: 1 1 do ln (do /di ) do 1 do 1 1 = + + + + Uo ho hod 2kw di hi di hid

(4.11)


4.2 ε- NTU

61

where Uo = ho = hi = hod = hi = kw = do = di =

the overall coefficient based on the outside area of the tubeW/m2 o C outside fluid film coefficient, W/m2 o C inside fluid film coefficient, W/m2 o C outside dirt coefficient (Fouling factor), W/m2 o C inside dirt coefficient,W/m2 o C thermal conductivity of the tube wall material, W/mo C tube outside diameter, m tube inside diameter, m

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

hd d µ Nu = = f Re, P r, , k L µw

!

(4.12)

and for natural convection as Ã

hd µ Nu = = f Gr, P r, k µw

!

(4.13)

Design equations for the heat transfer coefficient for various flow geometry (tube, plate) and configuration are given in Appendix 1. Design equation for the heat transfer coefficient for condensation and boiling is given also in appendix A. 4.1.5 Fouling factor (hid , hod ) Heat transfer may be degraded in time by corrosion, deposits of reaction products, organic growths, etc. These effects are accounted for quantitatively by fouling resistances. Extensive data on fouling factor are given TEMA standards. Typical fouling factors for common process and service fluids are given in the table 4.2. These values are for shell and tube heat exchangers with plain (not finned) tubes.

4.2

ε- NTU

The effectiveness (ε) of a heat exchanger is defined as the ratio between the actual heat load to the maximum possible heat load. ε=

Q Qmax

(4.14)

This is related to the heat exchanger size and capacity as ε = f (N T U, C)

(4.15)

Where N T U is number of transfer unit and is defined as NT U = N =

UA Cmin

(4.16)

and C is the heat capacity ratio defined using energy equation as: Q = Mh Cph (Thi − Tho ) = Mc Cpc (Tco − Tci )

(4.17)


62


Table 4.2. Fouling factor

Coefficient (W/m2 o C) Factor (resistance (m2 o C/W )

Fluid River water Sea water cooling water (towers) Towns water (soft) Towns water (hard) Steam condensate Steam oil free Steam oil traces Refrigerated brine Air and industrial gases Flue gases Organic vapor Organic liquids Light hydrocarbons Heavy hydrocarbons Boiling organics Condensing organics Heavy transfer fluids Aqueous salt solutions

3000-12000 1000-3000 3000-6000 3000-5000 1000-2000 1500-5000 4000-10000 2000-5000 3000-5000 5000-10000 2000-5000 5000 5000 5000 2000 2500 5000 5000 3000-5000

0.003-0.0001 0.001-0.0003 0.0003-0.00017 0.0003-0.0002 0.001-0.0005 0.00067-0.0002 0.0025-0.00001 0.0005-0.0002 0.0003-0.0002 0.0002-0.00001 0.0005-0.0002 0.0002 0.0002 0.0002 0.0005 0.0004 0.0002 0.0002 0.0003-0.0002

Mh Cph < Mc Cpc ⇒ Cmin = Mh Cph , Cmax = Mc Cpc

(4.18)

Mh Cpc > Mc Cpc ⇒ Cmin = Mc Cpc , Cmax = Mh Cph

(4.19)

Qmax = Cmin (Thi − Tci )

(4.20)

Cmin Cmax

(4.21)

Thi − Tho Tco − Tci , εc = Thi − Tci Thi − Tci

(4.22)

∆Tc Tspan

(4.23)

C=

εh =

ε=

where Tspan is defined in fig. 4.5 for counter current flow Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

4.2 ε- NTU

63

Thi

Tco Tho

Tspan

∆θ Tci 0

A

Figure 4.5. Temperature distribution in counter current flow

The ε equation for various heat exchanger configuration is given as • Parallel flow

1 − exp [−N (1 + C)] 1+C

(4.24)

1 − exp [−N (1 + C)] 1 − C exp [−N (1 − C)]

(4.25)

ε= • Counter current flow ε= • Cross flow 1. Both fluid unmixed mixed

"

exp(−N Cn) − 1 ε = 1 − exp Cn where

#

(4.26)

n = N −0.22

(4.27)

2. Both fluid mixed "

C 1 1 + − ε= 1 − exp(−N ) − 1 1 − exp(−N C) − 1 N

#−1

(4.28)

3. Cmax mixed, Cmin unmixed ε=

1 {1 − exp [−C (1 − exp(−N ))]} C

(4.29)


64


4. Cmax unmixed, Cmin mixed ½

¾

ε = 1 − exp −

1 [1 − exp(−N C)] C

(4.30)

• One shell pass, 2,4, 6 tube passes   

q

(1 + C 2 )

ε=2 1+C +  

• Condenser • Evaporator

h

q

i

h

q

i

−1  1 + exp −N (1 + C 2 ) 

 1 − exp −N (1 + C 2 ) 

(4.31)

ε = 1 − e−N

(4.32)

ε = 1 − e−N

(4.33)

Alternatively these equations are presented in a graphical form. The various curves of ε vs N T U can be found in textbooks like Kern (1964( and Perry and Green (2000).

4.3

Link between LMTD and NTU

• Cocurrent

µ

∆T1 ∆T2

¶

∆T1 ln ∆T2

¶

ln • Counter current

4.4

µ

µ

Thi − Tci Tho − Tco

¶

Thi − Tco = ln Tho − Tci

¶

= ln µ

= N h + Nc

(4.34)

= Nh − Nc

(4.35)

The Theta Method

Alternative method of representing the performance of heat exchangers may be given by Theta method [146] as Θ=

∆Tm Tspan

(4.36)

where ∆Tm is the mean temperature difference and Tspan is the maximum temperature difference (Thi −Tci ) (see Fig. 4.5). The Theta method is related is related to the associated ε and N T U methods by expressions Θ=

ε ∆Tm = Tspan NT U

(4.37)

The relationship between parameters are often presented in graphical form as shown in Fig.4.6. However, they all depend on finding ∆Tm or ∆Tlm


4.4 The Theta Method

65

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


66

5 Thermal Design

5

Thermal Design

5.1

Design Consideration

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

Liquids 1. Tube-side process fluids:1 to 2 m/s, maximum 4 m/s if required to reduce fouling: water 1.5 to 2.5 m/s 2. Shell side: 0.3 to 1/m/s Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

5.1 Design Consideration

67

Vapors For vapors, the velocity used will depend on the operating pressure and fluid density; the lower values in the range given below will apply to molecular weight materials Vacuum 50 to 70 m/s Atmospheric pressure 10 to 30 m/s High pressure 5 to 10 m/s 5.1.3 Stream temperature The closer the temperature approach used (the difference between the outlet temperature of one stream and the inlet temperature of the other stream) the larger will be the heat transfer area required for a given duty. The optimum value will depend on the application and can only be determined by making an economic analysis of alternative designs. As a general guide the greater temperature difference should be at least 20 o C. and the least temperature difference 5 to 7 o C for cooler using cooling water and 3 to 5 o C using refrigerated brine. The maximum temperature rise in recirculated cooling water is limited to around 30 o C. Care should be taken to ensure that cooling media temperatures are kept well above the freezing point of the process materials. When heat exchange is between process fluids for heat recovery the optimum approach temperatures will normally not be lower than 20 o C. 5.1.4 Pressure drop The value suggested below can be used as a general guide and will normally give designs that are near the optimum.

Liquids Viscosity<1 mN s/m2 ∆p< 35kN/m2 Viscosity=1 to 10mN s/m2 ∆p= 50-70 kN/m2

Gas and Vapors High vacuum Medium vacuum 1 to 2 bar Above 10 bar

0.4-0.8 kN/m2 0.1×absolute pressure 0.5×system gauge pressure 0.1×system gauge pressure

When a high-pressure drop is utilized, care must be taken to ensure that the resulting high fluid velocity does not cause erosion or flow -induced tube vibration. 5.1.5 Fluid physical properties In the correlation used to predict heat-transfer coefficients, the physical properties are usually evaluated at the mean stream temperature. This is satisfactory when the temperature change is small, but can cause a significant error when change in temperature is large. In these circumstances , a simple and safe procedure is to evaluate the heat transfer coefficients at the stream inlet and outlet temperatures and use the lowest of the Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

h

U2 (T1 −t2 ) U1 (T2 −t1 )

i

(5.1)

where U1 , U2 are evaluated at the end of the exchanger. If the variation is too large for these simple methods to be used it will be necessary to divide the temperature-enthalpy profile into sections and evaluate the heat transfer coefficients and area required for each section.

5.2

Design data

Before discussing actual thermal design, let us look at the data that must be furnished by the process licensor before design can begin: 1. flow rates of both streams. 2. inlet and outlet temperatures of both streams. 3. operating pressure of both streams. This is required for gases, especially if the gas density is not furnished; it is not really necessary for liquids, as their properties do not vary with pressure. 4. allowable pressure drop for both streams. This is a very important parameter for heat exchanger design. Generally, for liquids, a value of 0.5-0.7 kg/cm2 is permitted per shell. A higher pressure drop is usually warranted for viscous liquids, especially in the tubeside. For gases, the allowed value is generally 0.05-0.2 kg/cm2 , with 0.1 kg/cm2 being typical. 5. fouling resistance for both streams. If this is not furnished, the designer should adopt values specified in the TEMA standards or based on past experience. 6. physical properties of both streams. These include viscosity, thermal conductivity, density, and specific heat, preferably at both inlet and outlet temperatures. Viscosity data must be supplied at inlet and outlet temperatures, especially for liquids, since the variation with temperature may be considerable and is irregular (neither linear nor log-log). 7. heat duty. The duty specified should be consistent for both the shellside and the tubeside. 8. type of heat exchanger. If not furnished, the designer can choose this based upon the characteristics of the various types of construction described earlier. In fact, the designer is normally in a better position than the process engineer to do this. 9. line sizes. It is desirable to match nozzle sizes with line sizes to avoid expanders or reducers. However, sizing criteria for nozzles are usually more stringent than for lines, especially for the shellside inlet. Consequently, nozzle sizes must sometimes be one size (or even more in exceptional circumstances) larger than the corresponding line sizes, especially for small lines. 10. preferred tube size. Tube size is designated as O.D., thickness, length. Some plant owners have a preferred O.D., thickness (usually based upon inventory considerations), and the available plot area will determine the maximum tube length. Many plant owners prefer to standardize all three dimensions, again based upon inventory considerations. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

5.3 Tubeside design

69

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

12. materials of construction. If the tubes and shell are made of identical materials, all components should be of this material. Thus, only the shell and tube materials of construction need to be specified. However, if the shell and tubes are of different metallurgy, the materials of all principal components should be specified to avoid any ambiguity. The principal components are shell (and shell cover), tubes, channel (and channel cover), tubesheets, and baffles. Tubesheets may be lined or clad.

13. special considerations. These include cycling, upset conditions, alternative operating scenarios, and whether operation is continuous or intermittent.

5.3

Tubeside design

Tubeside calculations are quite straightforward, since tubeside flow represents a simple case of flow through a circular conduit. Heat-transfer coefficient and pressure drop both vary with tubeside velocity, the latter more strongly so. A good design will make the best use of the allowable pressure drop, as this will yield the highest heat-transfer coefficient. If all the tubeside fluid were to flow through all the tubes (one tube pass), it would lead to a certain velocity. Usually, this velocity is unacceptably low and therefore has to be increased. By incorporating pass partition plates (with appropriate gasketing) in the channels, the tubeside fluid is made to flow several times through a fraction of the total number of tubes. Thus, in a heat exchanger with 200 tubes and two passes, the fluid flows through 100 tubes at a time, and the velocity will be twice what it would be if there were only one pass. The number of tube passes is usually one, two, four, six, eight, and so on. 5.3.1 Heat-transfer coefficient The tubeside heat-transfer coefficient is a function of the Reynolds number, the Prandtl number, and the tube diameter. These can be broken down into the following fundamental parameters: physical properties (namely viscosity, thermal conductivity, and specific heat); tube diameter; and, very importantly, mass velocity. The variation in liquid viscosity is quite considerable; so, this physical property has the most dramatic effect on heat-transfer coefficient. The fundamental equation for turbulent heat-transfer inside tubes is: Ã a

N u = CRe P r

b

µ µw

!c

,

(5.2)

or k h=C D

Ã

GD µ

!a µ

Cp µ k

¶b Ã

µ µw

!c

(5.3)


70

where

5 Thermal Design

Nu = Pr = Re de A P u µw k Cp

hde k Cp µ k ρud µ 4A P

Nusselt number Prandtl number Reynolds number hydraulic diameter cross-sectional area wetted perimeter fluid velocity fluid viscosity at the tube wall temperature fluid thermal conductivity fluid specific heat

C=

    

0.021 gases 0.023 non-viscous liquid 0.027 viscous liquid a = 0.8 b = 0.3 for cooling b = 0.4 for heating c = 0.14

Viscosity influences the heat-transfer coefficient in two opposing ways- as a parameter of the Reynolds number, and as a parameter of Prandtl number. Thus, from Eq. 5.3: h ∝ µ0.8−0.33 = µ0.47

(5.4)

In other words, the heat-transfer coefficient is inversely proportional to viscosity to the 0.47 power. Similarly, the heat-transfer coefficient is directly proportional to thermal conductivity to the 0.67 power. These two facts lead to some interesting generalities about heat transfer. A high thermal conductivity promotes a high heat-transfer coefficient. Thus, cooling water (thermal conductivity of around 0.55 kcal/hm◦ C) has an extremely high heat-transfer coefficient of typically 6,000 kcal/hm2◦ C, followed by hydrocarbon liquids (thermal conductivity between 0.08 and 0.12 kcal/hm◦ C) at 250-1,300 kcal/hm2◦ C, and then hydrocarbon gases (thermal conductivity between 0.02 and 0.03 kcal/hm◦ C) at 50-500 kcal/hm2◦ C. Hydrogen is an unusual gas, because it has an exceptionally high thermal conductivity (greater than that of hydrocarbon liquids). Thus, its heat-transfer coefficient is toward the upper limit of the range for hydrocarbon liquids. The range of heat-transfer coefficients for hydrocarbon liquids is rather large due to the large variation in their viscosity, from less than 0.1 cP for ethylene and propylene to more than 1,000 cP or more for bitumen. The large variation in the heat-transfer coefficients of hydrocarbon gases is attributable to the large variation in operating pressure. As operating pressure rises, gas density increases. Pressure drop is directly proportional to the square of mass velocity and inversely proportional to density. Therefore, for the same pressure drop, a higher mass velocity can be maintained when the density is higher. This larger mass velocity translates into a higher heat-transfer coefficient. 5.3.2 Pressure drop The pressure drop due to friction exists because of the shear stress between the fluid and the tube wall. Estimation of the friction pressure drop is somewhat more complex and Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

5.3 Tubeside design

71

various approaches have been taken, for example the frictional pressure gradient is given as Ã ! 4τo dp 4f G2 = − = , (5.5) dz f d 2dρ where G is the mass flux in kg/m2 s and f is the friction factor calculated using a Blasiustype model as f=

 0.3164   Re0.25  

64 Re

Re

≥ 2320

Re < 2320 .

Integration of equation B.1 yields ∆p =

4f G2 L , 2ρ d

(5.6)

Mass velocity strongly influences the heat-transfer coefficient. For turbulent flow, the tubeside heat-transfer coefficient varies to the 0.8 power of tubeside mass velocity, whereas tubeside pressure drop varies to the square of mass velocity. Thus, with increasing mass velocity, pressure drop increases more rapidly than does the heat-transfer coefficient. Consequently, there will be an optimum mass velocity above which it will be wasteful to increase mass velocity further. Furthermore, very high velocities lead to erosion. However, the pressure drop limitation usually becomes controlling long before erosive velocities are attained. The minimum recommended liquid velocity inside tubes is 1.0 m/s, while the maximum is 2.5-3.0 m/s. Pressure drop is proportional to the square of velocity and the total length of travel. Thus, when the number of tube passes is increased for a given number of tubes and a given tubeside flow rate, the pressure drop rises to the cube of this increase. In actual practice, the rise is somewhat less because of lower friction factors at higher Reynolds numbers, so the exponent should be approximately 2.8 instead of 3. Tubeside pressure drop rises steeply with an increase in the number of tube passes. Consequently, it often happens that for a given number of tubes and two passes, the pressure drop is much lower than the allowable value, but with four passes it exceeds the allowable pressure drop. If in such circumstances a standard tube has to be employed, the designer may be forced to accept a rather low velocity. However, if the tube diameter and length may be varied, the allowable pressure drop can be better utilized and a higher tubeside velocity realized. The following tube diameters are usually used in the CPI: (1/4, 3/8, 1/2, 5/8, 3/4, 1, 1 1/4, 1 1/2 in. Of these, 3/4 in. and 1 in. are the most popular. Tubes smaller than 3/4 in. O.D. should not be used for fouling services. The use of small-diameter tubes, such as 1 in., is warranted only for small heat exchangers with heat-transfer areas less than 20-30 m2 . It is important to realize that the total pressure drop for a given stream must be met. The distribution of pressure drop in the various heat exchangers for a given stream in a particular circuit may be varied to obtain good heat transfer in all the heat exchangers. Consider a hot liquid stream flowing through several preheat exchangers. Normally, a pressure drop of 0.7 kg/cm2 per shell is permitted for liquid streams. If there are five such preheat exchangers, a total pressure drop of 3.5 kg/cm2 for the circuit would be permitted. If the pressure drop through two of these exchangers turns out to be only 0.8 kg/cm2 , the balance of 2.7 kg/cm2 would be available for the other three. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

72

5.4

5 Thermal Design

Shell side design

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

5.4 Shell side design

73

5.4.2 Tube layout patterns There are four tube layout patterns, as shown in Figure 5.1: triangular (30◦ ), rotated triangular (60◦ ), square (90◦ ), and rotated square (45◦ ).

Figure 5.1. Tubes layout pattern.

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

74

5 Thermal Design

For a triangular pattern, TEMA specifies a minimum tube pitch of 1.25 times the tube O.D. Thus, a 25- mm tube pitch is usually employed for 20-mm O.D. tubes. For square patterns, TEMA additionally recommends a minimum cleaning lane of 4 in. (or 6 mm) between adjacent tubes. Thus, the minimum tube pitch for square patterns is either 1.25 times the tube O.D. or the tube O.D. plus 6 mm, whichever is larger. For example, 20-mm tubes should be laid on a 26-mm (20 mm + 6 mm) square pitch, but 25-mm tubes should be laid on a 31.25-mm (25 mm ´ 1.25) square pitch. Designers prefer to employ the minimum recommended tube pitch, because it leads to the smallest shell diameter for a given number of tubes. However, in exceptional circumstances, the tube pitch may be increased to a higher value, for example, to reduce shellside pressure drop. This is particularly true in the case of a cross-flow shell. 5.4.4 Baffling Type of baffles. Baffles are used to support tubes, enable a desirable velocity to be maintained for the shellside fluid, and prevent failure of tubes due to flow-induced vibration. There are two types of baffles: plate and rod. Plate baffles may be single-segmental, double-segmental, or triple-segmental, as shown in Figure 5.2.

Figure 5.2. Types of baffles.

Baffle spacing. Baffle spacing is the centerline-to-centerline distance between adjacent baffles. It is the most vital parameter in STHE design. The TEMA standards specify the minimum baffle spacing as one-fifth of the shell inside diameter or 2 in., whichever is greater. Closer spacing will result in poor bundle penetration by the shellside fluid and difficulty in mechanically cleaning the outsides of the tubes. Furthermore, a low baffle spacing results in a poor stream distribution as will be explained later. The maximum baffle spacing is the shell inside diameter. Higher baffle spacing will lead to predominantly longitudinal flow, which is less efficient than cross-flow, and large unsupported tube spans, which will make the exchanger prone to tube failure due to flow-induced vibration. Optimum baffle spacing. For turbulent flow on the shellside (Re > 1,000), the heattransfer coefficient varies to the 0.6-0.7 power of velocity; however, pressure drop varies Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


75

to the 1.7-2.0 power. For laminar flow (Re < 100), the exponents are 0.33 for the heattransfer coefficient and 1.0 for pressure drop. Thus, as baffle spacing is reduced, pressure drop increases at a much faster rate than does the heat-transfer coefficient. This means that there will be an optimum ratio of baffle spacing to shell inside diameter that will result in the highest efficiency of conversion of pressure drop to heat transfer. This optimum ratio is normally between 0.3 and 0.6. Baffle cut. As shown in Figure 5.3, baffle cut is the height of the segment that is cut in each baffle to permit the shellside fluid to flow across the baffle. This is expressed as a percentage of the shell inside diameter. Although this, too, is an important parameter for STHE design, its effect is less profound than that of baffle spacing.

Figure 5.3. Baffle cut.

Baffle cut can vary between 15% and 45% of the shell inside diameter. Both very small and very large baffle cuts are detrimental to efficient heat transfer on the shellside due to large deviation from an ideal situation, as illustrated in Figure 5.4.

Figure 5.4. Effect of small and large baffle cuts.

It is strongly recommended that only baffle cuts between 20% and 35% be employed. Reducing baffle cut below 20% to increase the shellside heat-transfer coefficient or increasing the baffle cut beyond 35% to decrease the shellside pressure drop usually lead to poor designs. Other aspects of tube bundle geometry should be changed instead to achieve those goals. For example, doublesegmental baffles or a divided-flow shell, or even a cross-flow shell, may be used to reduce the shellside pressure drop. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

76

5 Thermal Design

For single-phase fluids on the shellside, a horizontal baffle cut (Figure 5.5) is recommended, because this minimizes accumulation of deposits at the bottom of the shell and also prevents stratification. However, in the case of a two-pass shell (TEMA F), a vertical cut is preferred for ease of fabrication and bundle assembly.

Figure 5.5. Baffle cut orientation

5.4.5 Equalize cross-flow and window velocities Flow across tubes is referred to as cross-flow, whereas flow through the window area (that is, through the baffle cut area) is referred to as window flow. The window velocity and the cross-flow velocity should be as close as possible - preferably within 20% of each other. If they differ by more than that, repeated acceleration and deceleration take place along the length of the tube bundle, resulting in inefficient conversion of pressure drop to heat transfer. 5.4.6 Shellside stream analysis (Flow pattern) On the shellside, there is not just one stream, but a main cross-flow stream and four leakage or bypass streams, as illustrated in Figure 5.6. Tinker (4) proposed calling these streams the main cross-flow stream (B), a tube-to-baffle-hole leakage stream (A), a bundle bypass stream (C), a pass-partition bypass stream (F), and a baffle-to-shell leakage stream (E). While the B (main cross-flow) stream is highly effective for heat transfer, the other streams are not as effective. The A stream is fairly efficient, because the shellside fluid is in contact with the tubes. Similarly, the C stream is in contact with the peripheral tubes around the bundle, and the F stream is in contact with the tubes along the passpartition lanes. Consequently, these streams also experience heat transfer, although at a lower efficiency than the B stream. However, since the E stream flows along the shell wall, where there are no tubes, it encounters no heat transfer at all. The fractions of the total flow represented by these five streams can be determined for a particular set of exchanger geometry and shellside flow conditions by any sophisticated heatexchanger thermal design software. Essentially, the five streams are in parallel and flow along paths of varying hydraulic resistances. Thus, the flow fractions will be such that the pressure drop of each stream is identical, since all the streams begin and end at the inlet and outlet nozzles. Subsequently, based upon the efficiency of each of these streams, the overall shellside stream efficiency and thus the shellside heat-transfer coefficient is established. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


77

Figure 5.6. Tube arrangement

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

78

5 Thermal Design

Besides these methods there is some proprietary methods putout by various organization for use by their member companies. A number of these method are based on one of the above methods. Some are based upon a judicious combination of methods 3 and 4 above and supplemented by further research data. Among the most popular of the proprietary methods, judged by their large clientele are • Heat Transfer Research Inc. (HTRI), Alliambra, california. This method is also known as stream analysis method. • Heat Transfer and Fluid Flow Service (HTFS), Engineering Science Division, AERE, Harwell, United Kingdom Method. In this work only Kern’s method is given below. Bell-Delaware method may be found in Coulson and Richardson’s 5.4.8

Heat transfer coefficient Ã

N u = 0.36Re where

Nu = Pr = Re = de = A= P = G= As = pt = Ds = lB =

hde k Cp µ k Gde µ 4A P

M As (pt −do )Ds lB pt

0.55

Pr

1/3

µ µw

!0.14

,

(5.7)

Nusselt number Prandtl number Reynolds number hydraulic diameter cross-sectional flow area wetted perimeter Mass flux fluid viscosity at the tube wall temperature pitch diameter shell diameter Baffle spacing

Hydraulic diameter (Fig. 5.1) de =

5.4.9

      

p2t −πd2o /4 πdo

for square pitch

0.87p2t /2−πd2o /8 πdo /2

for equilateral triangular pitch

Pressure drop µ

Ds ∆p = 4f d where

¶Ã

ρu2 2

 0.3164   Re0.25

f = 

64 Re

!µ

Re

L lb

¶Ã

µ µw

!−0.14

,

(5.8)

≥ 2320

Re < 2320 .

L=tube length lB = baffle spacing. The term (L/lB ) is the number of times the flow crosses the tube bundle=(NB + 1). Where NB is the number of baffles. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

5.5 Design Algorithm

5.5

79

Design Algorithm

Step1 Specification Define duty Q Make energy balance if needed to calcualted unspecified flow rates or temperature Q=Mccpc(Tc2-Tc1)=MhCph(Th1-Th2)

Step 9 Estimate tube-side heat transfer coefficient Step 10 Decide baffle spacing and estimate shell side heat transfer coefficient

Step2 Calculate physical properties

Step 11 Calculate overall heat transfer Coefficient including fouling factors Uo,cal

Step3 Assume value of overall coefficient Uo,ass Step 4 Decide number of shell and tube passes Calculate ∆Tlm, F and ∆Tm

No Set Uo,ass=Uo,cal

0<(Uo,cal-Uo,ass)/Uo,ass<30

Step 12 Estimate tube and shell side pressure drop

Step 5 Determine heat transfer area required Ao=q/Uo,ass∆Tm Step 6 Decide type, tube size, material, layout Assign fluids to shell or tube

No

Step 13 Estimate cost of heat exchanger

Step 7 Calculate number of tubes Step 8 Calculate shell diameter

Is pressure drops within specification?

Yes

Can design be optmized to reduce cost? Accept design

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


80

6 Specification sheet

6

Specification sheet

Specification sheet is a data sheet that contains the information provided by the customer to the vendor for the process and mechanical designs of an exchanger. After the process design is done, the engineer fills in some further information. The rest of the information is filled after the mechanical design is completed. The specification sheet is a medium of communication between different parties involved in the procurement, design and fabrication of heat exchanger. It is also used to compare the performance of the installed unit with the design conditions.

6.1

Information included

The information contained in the sheet is best decribed by a data sheet. Although each company has its own version of data sheet, the most popular one is that of the TEMA standards. It is similar to that of API standard 660. It contains the fluid • flow rate and properties, • heat duty, • heat transfer coefficient, • fouling resistance, • details about the shell and tube size, • materials, • baffle nozzle, etc.. Some variations include information for alternate designs and different systems of units (British, SI, metric).

6.2

Information not included

The regarding the type of flanges, studs, vent and relief valves, drains lines, welding, inspection and testing requirement of the material of construction, etc.. are not given in the specification sheet.

6.3

Operation conditions

The following operating conditions regarding the exchanger operation should be known to the thermal designer for critical application. 1. Start-up condition and procedure 2. Normal operating conditions 3. Upset and emergency conditions 4. shut down conditions and procedure 5. possibility of switching the shell-side and tube tube side fluid for better design 6. possibility of increasing the allowable pressure drop to control the fouling 7. beside these the spec-sheet should provided with other information concerning the composition of the streams, their thermal and physical properties and any phase change occurring. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

6.4 Bid evaluation

6.4

81

Bid evaluation

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


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6 Specification sheet

Figure 6.1. Data sheet


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7

Storage, Installation, Operation and Maintenance

Proper storage, installation handling and correct start up emergency, and shutdown procedure are important for the successful working of a well designed and fabricated heat exchanger. regular cleaning, maintenance and repairs are necessary to ensure trouble free operation of the unit for its designed life span. These will be discussed in the following sections. NOTE: Before placing your equipment in operation, environment and service conditions should be checked for compatibility with materials of construction. Contact your nearest heat exchanger Standard representative if you are not sure what the actual materials of construction are. Successful performance of heat transfer equipment, length of service and freedom from operating difficulties are largely dependent upon: 1. Proper thermal design. 2. Proper physical design. 3. Storage practice prior to installation. 4. Manner of installation, including design of foundation and piping. 5. The method of operation. 6. The thoroughness and frequency of cleaning. 7. The materials, workmanship, and tools used in maintenance and making repairs and replacements. Failure to perform properly may be due to one or more of the following: 1. Exchanger being dirty. 2. Failure to remove preservation materials after storage. 3. Operating conditions being different than design conditions. 4. Air or gas binding. 5. Incorrect piping connections. 6. Excessive clearances between internal parts due to corrosion. 7. Improper application.

7.1

Storage

Standard heat exchangers are protected against the elements during shipment. If they cannot be installed and put into operation immediately upon receipt at the jobsite, certain precautions are necessary to prevent deterioration during storage. Responsibility for integrity of the heat exchangers must be assumed by the user. The manufacturer will not be responsible for damage, corrosion or other deterioration of heat exchanger equipment during transit and storage. Good storage practices are important, considering the high costs of repair or replacement, and the possible delays for items which require long lead times for manufacture. The following suggested practices are provided solely as a convenience to the user, who shall make his own decision on whether to use all or any of them. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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7 Storage, Installation, Operation and Maintenance

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

Date Inspector’s name Identification of unit or item Location Condition of paint or coating Condition of interior Is free moisture present? Has dirt accumulated? Corrective steps taken

9. To locate ruptured or corroded tubes or leaking joints between tubes and tubesheets, the following procedure is recommended: Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

7.2

Installation

7.2.1 Installation Planning 1. On removable bundle heat exchangers, provide sufficient clearance at the stationary end to permit the removal of the tube bundle from the shell. On the floating head end, provide space to permit removal of the shell cover and floating head cover. 2. On fixed bundle heat exchangers, provide sufficient clearance at one end to permit removal and replacement of tubes and at the other end provide sufficient clearance to permit tube rolling. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

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

7.3

Operation

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

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

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

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


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8

Heat exchanger tube side mainenance (Repair vs replacement

(This subject of chapter is collected from: Bruce W Schafer Framatome ANP, Inc. 155 Mill Ridge Road Lynchburg, VA 24502 (434) 832-3360 [email protected]) Abstract The traditional method of repairing degraded tubes in shell-and-tube heat exchangers is to remove the effected tubes from service by plugging. Since heat exchangers are designed with excess heat transfer capability, approximately 10% of tubes can be plugged before performance is affected. When the number of plugged tubes becomes excessive, heat exchanger efficiency is lost, resulting in reduced power output, high system pressure drop, further heat exchanger damage, or abnormal loads placed on other plant heat exchangers. As an option to component retubing or replacement, repair methods, including tube sleeving and tube expansion, have proven to be an effective method to repair defective tubes and keep the existing heat exchanger in service. For the sleeving process, a new tube section is installed inside the existing tube to bridge across the degraded area. Tube expansion is used to close off a gap between the tube and the tubesheet or end plate (to eliminate a leak path) or between the tube and tube support (to minimize vibration). While not all heat exchangers can be returned to their original design condition by performing tube repairs, in some instances it may be possible to get many more years of useful life out of a heat exchanger at a fraction the cost of replacement. This paper presents options which the Plant Maintenance Engineer should consider in making the repair versus replacement decision. This includes the repair options (sleeving and tube expansion), other conditions within the heat exchanger, and the effect of tube repair on heat exchanger performance.

8.1

Introduction

Traditionally, when maintenance is performed on shell-and-tube heat exchangers, the only options considered when tube defects are found are to plug tubes and, when the number of plugs became too great, replace the heat exchanger. The decision to replace the heat exchanger was based on a number of factors. These included: the number of tubes plugged, the number of forced outages due to tube damage (and the cost associated with replacing lost power and repairing the damaged tubes), the impact that the plugged heat exchanger is having on the plant (due to lost flow or heat transfer surface area), the rate at which tube plugging is occurring, the availability of funds to replace the heat exchanger, and the expected life of the unit (how much longer will the unit operate before retirement). From a sampling of industry data, tube failures have been shown to cause between 31% to 87% (depending on the data source) of the events related to feedwater heaters (1). Since so many of the failures were related to the tubing, the replacement of an entire heat exchanger due to damage in one area is an expensive as well as a schedule and manpower intensive option. The typical means for major heat exchanger repair included complete replacement, rebundling, and retubing, as described below. • For the replacement option, the entire heat exchanger shell and tube bundle are replaced with a new unit. • For rebundling, the shell is temporarily removed from the heat exchanger and the old tube bundle, including, at a minimum, tubes, tube supports, and tubesheet, are removed. A new tube bundle is inserted and the shell is welded back in place. • For retubing, either the shell (u-tube design) or tube side access cover (straight Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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8 Heat exchanger tube side mainenance (Repair vs replacement

tubes) is removed from the heat exchanger and the old tubes are removed from the bundle. New tubes are then inserted and re-attached to the tubesheet (typically by either mechanical expansion, welding, or both). In many instances, the existing shell side hardware is used as-is, although some modifications may be made. Retubing is typically performed on straight tube heat exchangers, such as condensers and coolers. Since the 1970’s, tube sleeving has been used to allow damaged tubes to remain in service. The sleeves are installed by various means (roll, explosive, or hydraulic expansion, explosively welded, or press-fit or epoxied in place) over the defective area of the tube. Through the use of sleeving, which is a low-cost option to retubing, rebundling, or replacement, the useful life of a heat exchanger can be economically extended. The decision to perform sleeving also can be made with short notice as opposed to replacement (2-6 weeks compared with 18 months), possibly allowing repairs to be performed the same outage that the damage is noted. Tube expansion also can be performed to minimize or eliminate leakage within heat exchangers. In the tubesheet, tubes can be re-expanded to strengthen the original tube-to-tubesheet joint, reducing or eliminating leakage and prolonging the life of the heat exchanger. Expansions also can be made deep within the tube to expand the tube into tube support plates and end plates. These expansion can reduce tube-to-plate clearance for vibration control or, at end plates, to minimize steam flow from the high to low pressure side of the plate.Since the 1970’s, tube sleeving has been used to allow damaged tubes to remain in service. The sleeves are installed by various means (roll, explosive, or hydraulic expansion, explosively welded, or press-fit or epoxied in place) over the defective area of the tube. Through the use of sleeving, which is a lowcost option to retubing, rebundling, or replacement, the useful life of a heat exchanger can be economically extended. The decision to perform sleeving also can be made with short notice as opposed to replacement (2-6 weeks compared with 18 months), possibly allowing repairs to be performed the same outage that the damage is noted. Tube expansion also can be performed to minimize or eliminate leakage within heat exchangers. In the tubesheet, tubes can be re-expanded to strengthen the original tubeto-tubesheet joint, reducing or eliminating leakage and prolonging the life of the heat exchanger. Expansions also can be made deep within the tube to expand the tube into tube support plates and end plates. These expansion can reduce tube-to-plate clearance for vibration control or, at end plates, to minimize steam flow from the high to low pressure side of the plate.

8.2

Repair vs. Replace - Factors To Consider

There are numerous factors to consider when deciding whether to repair the tubes in a heat exchanger or to perform a larger repair scope and rebundle or replace the component. The following factors should be considered when making the repair vs. replace decision. 1. The budget available for repair or replacement needs to be determined. Typically, the cost of performing a substantial heat exchanger repair (consisting of plug removal, tube inspection, tube expansion, and sleeving) is less than 10% of the cost of replacing the unit. Because of the lower cost, the payback time on the repair option is much shorter than for replacement. If the heat exchanger is critical to plant operation (either from a safety, efficiency, or power production standpoint) or is resulting in costly forced outages, it may be possible to justify a 3 repair to the unit in the near-term and a scheduled replacement when a longer outage can be planned. If there are a large number of tube plugs to remove, or if they are difficult to remove (explosive or welded), then the cost to repair the heat exchanger will increase, and the scheduled time needed on-site may not fit within the outage window. If it appears that tube repair may be possible, it may be worthwhile to plug tubes, using removable plugs, until a certain quantity Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

8.3

Heat Exchanger maintenance Options

There have always been options available to either repair or replace heat exchanger tubes in the event that tube leakage or degradation is present. The repair options include: 1. Plug the tube Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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2. Sleeving 3. Tube expansion The replacement options include 1. Retubing 2. Rebundling 3. Replace with new unit

8.4

Repair option

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

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

Figure 8.1. Heat Exchanger Sleeve Designs


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Full Length Sleeve These sleeves are installed from one end of the tube to the other in straight tubed heat exchangers. After insertion, the full length of the sleeve is expanded into the parent tube. This step serves the dual purpose of maintaining heat transfer as high as possible (typically 75%-90%) while minimizing flow pressure drop through the tube. After the full length expansion step, shown in Figure 8.2, the sleeve ends are trimmed flush with the existing tube ends and the sleeve is roll expanded into the tubesheet. The full length sleeve is typically used in a condenser or cooling water heat exchanger when the tubes have multiple defects along their length. Full length sleeving is an attractive option when a relatively small percentage of the tubes require repair. Through sleeving, the majority of the tube heat transfer area can be left in service, resulting in a heat exchanger that is close to its as designed condition. Full length sleeving is comparable in many ways to retubing in the methods employed to install the sleeves. However, since removal of the existing tube is not required, and the typical number of tubes that will be full length sleeved are below the number that would be retubed, the cost for material and manhours are much less than for retubing, making sleeving a cost-effective option to return and keep tubes in service.

Figure 8.2. Full Length Sleeve Expansion

Partial Length Structural Sleeve This type of sleeve is used to repair shorter defects in the tube. The sleeve can be installed anywhere along the straight length of the tube. Various methods are used to expand the sleeve in place. These include roll expansion (both in the tubesheet and in the freespan portion of the tube), hydraulic expansion in the freespan portion of the tube, and full length expansion. These expansion types are discussed below. The installation of a hydraulically expanded sleeve is shown in Figure 8.3. • If one end of the sleeve is in the tubesheet, a torque-controlled roll expansion will be made. This expansion is similar to the original tube-to-tubesheet roll. Freespan roll expansions are made to either a torque controlled setting or to a diameter controlled hardstop setting. Usually, freespan roll expansions are only used when the sleeve length is relatively short, since it can be difficult to insert a roll expander deep into the tube. Both the tubesheet and freespan roll expansion parameters are set so that they can provide both the structural and leakage requirements for the sleeve. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

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Figure 8.3. Partial Length - Hydraulically Expanded Structural Sleeve Installation

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

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

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– if the tube-to-sleeve crevice needs to be eliminated in a hostile water chemistry environment – if there are large eddy current probe fill factor restrictions

Figure 8.4. Required Clearance for Sleeve Installation

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

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Figure 8.5. Required Clearance for Sleeve Installation

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

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

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If it appears that tube expansion is possible, then information will be needed to ensure that the heat exchanger is properly repaired. The following information is used when planning for tube expansion. • Tube expansion will need to be coordinated with eddy current inspection and plug removal. • The tube expander design (diameter and length) needs to be based on the requirements for the expansion. Before performing tube expansions into heat exchanger tubes, testing needs to be performed to set the tooling operating parameters. Depending on the type of expansion, these tests may include setting the rolling torque for tubesheet re-expansions or setting the hydraulic expansion constants for end plate expansions. In addition, for the tube-intotubesheet re-expansion process, qualification testing should be performed. This would consist of hydrostatic leak and pressure tests and temperature and pressure cycling. These tests would verify that the expansion parameters were set correctly for the tube reexpansions. exchanger.

8.5

Replacement option

8.5.1 Retubing The tubes can be replaced, if the unit has: • straight tubes, • good access, and • the remaining components (shell, tube supports, internal structural pieces) of the heat exchanger are in good shape. The old tubes are removed from the unit and new ones, typically manufactured from an improved material, are inserted, and then expanded, into place. Insertion of the new tubes is shown in Figure 8.6. In addition to performing retubing to replace damaged tubes, retubing has been performed to eliminate detrimental materials (such as copper from condenser tubes) to minimize damage to other equipment within the plant (nuclear steam generators or fossil boilers).

Figure 8.6. Condenser Retubing


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8.5.2 Rebundling Some heat exchangers are designed to be rebundled rather than replaced. For these units the entire tube bundle, including tubes, tubesheet, and tube supports are replaced, as shown in Figure 8.7. The original shell and any other internal structural pieces would be reused (although any necessary internal repairs could be made when the shell was removed). The new tube bundle can be manufactured to ensure that original design problems with the existing unit are corrected. However, the same basic design must be maintained since the new bundle must fit within the existing heat exchanger shell. Rebundling costs about 15-25% more than retubing.

Figure 8.7. Heat Exchanger Rebundling

8.5.3 Complete replacement (New unit) A third and typically widely used option is to replace the entire heat exchanger, as shown in Fig.8.8 . Full replacement allows alternate tube materials, changes in heat transfer area, and structural changes to be employed, including added clearances in areas where erosion or other problems may be occurring, to ensure that the current heat exchanger problems do not re-occur in the future. However, the cost associated with a full replacement is the greatest of the three options, about 5% more than for rebundling . In addition, there are no guarantees that the new heat exchanger design will not have new, unanticipated problems.

Figure 8.8. Heat Exchanger Replacement


8.6 Conclusions

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Conclusions

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


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9

9 Troubleshooting

Troubleshooting

9.1

Heat exchangers’ problems

Heat transfer equipment provides the economic and process viability for many plant operations. The basis for successful application of such equipment depends on the designer. The problem that should be anticipated by the design to avoid high maintenance or cleaning and costly shut down production include: 1. Fouling 2. Leakage 3. Corrosion To anticipate maintenance problems the designer should need to be familiar with the plant location, process flow sheet, plant operation. Some of the questions that must be considered are: 1. will the heat exchanger need cleaning? how often? what cleaning method will be used? 2. what penalty will the plant pay for leakages between the tubeside and shell side? 3. what kind of production upsets can occur that could affect the heat exchanger?will cycling occur? 4. how will heat exchanger be started up and shut down? 5. will the heat exchanger be likely to require repairs? if so, will the repairs present any special problem?

9.2 9.2.1

Fouling 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 Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

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3. surface structure • Fluid Properties 1. 2. 3. 4.

suspended solids dissolved solids dissolved gases trace elements

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

9.3 Leakage/Rupture of the Heat Transfer Surface

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Leakage/Rupture of the Heat Transfer Surface

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

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9.4

9 Troubleshooting

Corrosion

The heat transfer surface reacts chemically with elements of the fluid stream producing a less conductive, corrosion layer on all or part of the surface. 9.4.1

Corrosion effects

1. Premature metal failures 2. the deposit of corrosion products reduce both heat transfer and flow rate. 9.4.2 Causes of corrosion High content of total dissolved solids (TDS), the dissimilarity of the metal, dissolved oxygen, penetrating ions like chlorides and sulphates, the low pH and presence of various other impurities are the prime cause of corrosion in the heat exchanger. 9.4.3

Type of corrosion

• stress corrosion • galvanic corrosion • uniform corrosion • Pitting • Crevice Corrosion 9.4.4

Stress corrosion

• Differential expansion between tubes and shell in fixed-tube-sheet exchangers can develop stresses, which lead to stress corrosion. • Overthinning: Expanding the tube into the tube sheet reduces the tube wall thickness and work-hardens the metal. • The induced stresses can lead to stress corrosion. Controlling Stress Corrosion Cracking • Proper selection of the appropriate material. • Remove the chemical species that promotes cracking. • Change the manufacturing process or design to reduce the tensile stresses. 9.4.5 Galvanic corrosion Galvanic corrosion is frequently referred to as dissimilar metal corrosion. Galvanic corrosion can occur when two dissimilar materials are coupled in a corrosive electrolyte. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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

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9.4.9 Materials of Construction The various parts of the heat exchanger (tube, shell, tube sheet, baffles, front head, rear head, nozzles,...) may be manufactured from same metal or dissimilar metals. Individual components may be fabricated from single metal or bimetallic. For the selection of material of construction, the corrosion chart must be consulted (Appendix C of Coulson and Richardson [29]). The chart gives metal (alloy) vs chemical at various temperatures. Note:Before using the corrosion chart the notation given should read thoroughly. 9.4.10 Fabrication Expanding the tube into the tube sheet reduces the tube wall thickness and work-hardens the metal. The induced stresses can lead to stress corrosion. Differential expansion between tubes and shell in fixed-tube-sheet exchangers can develop stresses, which lead to stress corrosion. When austenitic stainless-steel tubes are used for corrosion resistance, a close fit between the tube and the tube hole is recommended in order to minimize work hardening and the resulting loss of corrosion resistance. In order to facilitate removal and replacement of tubes it is customary to roller-expand the tubes to within 3 mm of the shellside face of the tube sheet. A 3-mm- long gap is thus created between the tube and the tube hole at this tube-sheet face. In some services this gap has been found to be a focal point for corrosion. It is standard practice to provide a chamfer at the inside edges of tube holes in tube sheets to prevent cutting of the tubes and to remove burrs produced by drilling or reaming the tube sheet. In the lower tube sheet of vertical units this chamfer serves as a pocket to collect material, dirt, etc., and to serve as a corrosion center. Adequate venting of exchangers is required both for proper operation and to reduce corrosion. Improper venting of the water side of exchangers can cause alternate wetting and drying and accompanying chloride concentration, which is particularly destructive to the series 300 stainless steels. Certain corrosive conditions require that special consideration be given to complete drainage when the unit is taken out of service. Particular consideration is required for the upper surfaces of tube sheets in vertical heat exchangers, for sagging tubes, and for shell-side baffles in horizontal units.

9.5

Troubleshooting

This chapter presents potential failure mechanisms for heat transfer equipment and suggests design alternatives for reducing the risks associated with such failures. The types of heat exchangers covered in this chapter include: • Shell and tube exchangers • Air cooled exchangers • Direct contact exchangers • Others types including helical, spiral, plate and frame, and carbon block exchangers This chapter presents only those failure modes that are unique to heat transfer equipment. Some of the generic failure scenarios pertaining to vessels may also be applicable to heat transfer equipment. Unless specifically noted, the failure scenarios apply to more than one class of heat transfer equipment. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

9.6 Past failure incidents

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113

Past failure incidents

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

114

9 Troubleshooting

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

9.7

Failure scenarios and design solutions

Table 9.1 presents information on equipment failure scenarios and associated design solutions specific to heat transfer equipment.


9.7 Failure scenarios and design solutions

115

Figure 9.1. troubleshooting



116

9 Troubleshooting



9.8

Discussion

9.8.1 Use of Potential Design Solutions Table To arrive at the optimal design solution for a given application, use Tables 9.1-9.4 in conjunction with the design basis selection methodology presented earlier. Use of the design solutions presented in the table should be combined with sound engineering judgment and consideration of all relevant factors. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

9.8 Discussion

117

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

118

9 Troubleshooting

mitigation measures, such as siting outside the potential fire zone or diking with sloped drainage, may be used to reduce the likelihood and magnitude of external fire impinging on the heat exchanger. Alternative heat exchanger designs may also be used to reduce the surface area presented to an external fire.

9.9

Troubleshooting Examples

9.9.1

Shell side temperature uncontrolled

55 oC

oC

55-62 uncontrolled

67 oC

Control vlave

55-62 oC controlled

Bypass

Control vlave

30 oC Organic o 125 C Water Symptom: Shellside outlet temperaturee cannot be controlled within desired range (55-62 oC) by controlling flow of 125 oC water to tubes. The heat exchanger is 4 tube pass.

30 oC Organic o 70 C Water Diagnosis: Heat exchanger is Cure: Tube side water considerablyo versized for the temperature reduced to 70oC duty (because of an alternative and control valve removed. service). Temperature correction Control valve is installed factors F for LMTD fluctuate in new shellside bypass line widely with small changes in tube side flow

Figure 9.5. Shell side temperature uncontrolled

9.9.2

Shell assumed banana-shape 560 oC

200 oC

487 oC

belows joint

Symptom: Shell assumed banana shape and piping 600 oC connections leaked. leakage between tube and shell side Diagnosis: vertically cut baffle and inlets and outlets of top shell side, caused stratification of gases at top of shell. Poor distribution of hot gases lead to unequal expansionof tubes Cure: increase the number of baffles from two to three; weld baffles in the shell; install sealing strips at edges of bundle; installed three concentric cones in tube side inlet; install vapor belt - for shellside inlet nozzle; change baffles from vertical to horizontal cut.

Figure 9.6. Shell assumed banana-shape


9.9 Troubleshooting Examples

9.9.3

119

Steam condenser performing below design capacity Vent

Steam

ondensate

Symptom: Air cooled steam condensor performing below design capacity. Diagnosis: Careful measurement tube levels discloded that tubes sloped 1/4 inch in wrong direction (rising toward condensate end) Cure: Raise inlet end to obtain 2 inch slope toward condensate outlet

Figure 9.7. Steam condenser performing below design capacity

9.9.4 Steam heat exchanger flooded When a heat exchanger ”stalls,” condensate floods the steam space and causes a variety of problems within the exchanger:

Figure 9.8. Conventional motor driven condensate pump system

• Control hunting: As condensate backs up in the exchanger, the heat transfer rate to the process is greatly reduced. The control valve opens wide enough to allow flow into the exchanger. As condensate drains out, the steam space is now greater and the steam pressure increases. The process overheats, the control valve closes down, and the cycle repeats. • Temperature shock: Condensate backed up inside the steam space cools the tubes that carry the process fluid. When this sub-cooled condensate is suddenly replaced by hot steam due to poor steam trap operations, the expansion and contraction of the tubes stress the tube joints. Constantly repeating this cycle causes premature failure. • Corrosion from: 1. Flooding - A flooded heat exchanger will permit the oxygen to dissolve, as well as carbon dioxide and other gases found in the steam. Because the condensate Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

120

10 Unresolved problems in the heat exchangers design

is often sub-cooled due to the time it is in the exchanger, these gases are more readily dissolved. Together the cool condensate and dissolved gases are extremely corrosive and will tend to decrease the efficiency of the heat exchanger and reduce the heat transfer through the tubes. 2. Steam collapse - Under very low loads with the steam valve closed, the steam volume collapses to smaller volume condensate, inducing a vacuum. When the vacuum breaker opens, atmospheric air and condensate mix inside the exchanger, increasing the possibility of corrosion of the tubes, shells, tube sheet and tube supports. 3. Freezing - Steam/air coils cannot afford poor condensate drainage, especially if the coil experiences air below freezing temperature. Condensate backed up inside the coil will freeze, often within seconds, depending on the air temperature. A low temperature detection thermostat is recommended on the coil leaving side to sense freezing conditions. As we previously explained, the only way to avoid ”stall” is to eliminate back pressure on the steam trap. There are a number of options available for designing a system that greatly reduces the risk of ”stall.” The following are two such options: • Install the heat exchanger in a position so that the condensate freely drains by gravity to the condensate return line. In many cases this is not possible because of existing piping around the area in which the heat exchanger is needed (e.g., the heat exchanger is installed at a level lower than the condensate return tank). • Use an electric or pressure driven condensate pump package installed below the steam trap to pump condensate back to the boiler. In actual practice, the first option may not be possible, and so the use of electric or pressure driven pumps to return condensate to the boiler room should be considered.

10

Unresolved problems in the heat exchangers design

1. Accurate data on the thermodynamic properties: These are needed for both pure fluid and mixtures in single phase and two phase system under extremes conditions. It would be best if more predictable methods could be obtained 2. fouling (predictive method not available) 3. flow induced vibration (prediction) 4. two phase flow (flow regime) 5. boiling of mixture (heat transfer coefficient) 6. turbulence (better understanding)

10.1

Future trend

1. Stepwise calculation of overall heat transfer coefficient instead of assumption 2. Thermodynamic properties from built-in subroutines 3. workshops fabrication drawings. 4. better transportation facilities for the shell of heat exchanger. 5. computer design code Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

Bibliography

121

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[140] Steiner, D.: Strömungssieden Gesättiger Fl¨ ussigkeiten. in: VDI, VDI-GVC: VDIWärmeatlas. 8. Aufl., Springer-Verlag, Berlin, 1997, Abschnitt, Dbb [141] Stephan, K.: Two-phase heat exchange for new refrigerants and their mixtures. Int. J. Refrig., Vol. 18, No. 3, 1995, 198-209 [142] Stephan, K.; Korner, M.: Berechnung des Warme¨ ubergangs verdampfender binärer Fl¨ ußigkeitsgemische. Chem. Ing. Tech., Vol. 2, 1969, 161-169 [143] Stephan, K.; Preusser, P.: Heat transfer and critical heat flux in pool boiling of binary and ternary mixtures. German Chem. engng, Vol. 2, 1979, 161-169 [144] Storek, H.; Brauer, H.: Reibungsdruckverlust der adiabaten Gas/Fl¨ ußigkeitsstömung in horizontalen und vertikalen Rohren. VDI-Forschungsheft, Nr. 599, VDIVerlag GmbH, D¨ usseldorf, Germany, 1980 [145] Taitel, Y.; Dukler, A.E.: A model for predicting flow regime transitions in horizontal and near horizontal Gas-Liquid flow. AICHE J., Vol. 22, No. 2, 1985, 43-55 [146] In: Heat exchanger design handbook, New York: Hemisphere, Volo. 1, section 1.5 [147] Tubular Exchanger Manufacturers Association, ”Standards of the Tubular Exchanger Manufacturers Association,” 7th ed., TEMA, New York (1988). [148] Thome, J. R.: Prediction of binary mixture boiling heat transfer coefficient using only phase equilibrium data. Int. J. of Heat and Mass Transfer, Vol. 26, 1983, 965974 [149] Thome, J. R.: Enhanced boiling of mixture. Chemical Engineering Service, Vol. 42, No. 8, 1986, 1909-1917 [150] Tillner-Roth, R.: Die thermodynamischen Eigenschaften von R152a, R134a und ihren Gemischen-Messungen und Fundamental Gleichungen. Dissertation, Universität Hannover; Germany, 1993 [151] Tillner-Roth, R.: Fundamental equations of state. Shaker-Verlag, Aachen, Germany, 1998 [152] Tillner-Roth, R.; Baehr, H.D.: An international standard formulation of the thermodynamic properties of 1,1,1,2-tetraflouroethane (HFC-134a) covering temperatures from 170 K to 455 K at pressure up to 70 MPa. J. Phys. Chem. Ref. data., Vol. 23, 1994, 657 [153] Tillner-Roth, R.; Li, J.; Yokozenki, A.; Sato, H.; Watanabe, K.: Thermodynamic properties of pure and blended Hydroflourocarbons (HFCs) Refrigerants. Japan Society of Refrigeration and Air Conditioning Engineers, 1997 [154] Urso, M.E.; Wadekar, V.V; Hewitt, G.F.: Heat transfer at the dryout and near dryout regions in flow boiling. 12. Proc. of the 12. Int. Heat Tranasfer Conference, Grenoble, Vol.3. 2002, 701-706. [155] van Wijk, W.R.; Vos, A.S.; van Stralen, S.J.D.: Heat transfer to boiling binary liquid mixtures. Chem. Engng. Sci., Vol. 5, 1956, 68-80 [156] Verma, H.K.; Sharma, C.P.; Mishra, M.P.: Heat transfer coefficients during forced convective evaporation of R12 and R22 mixtures in annular flow regime. Proc. XV Int. Congr. Refrig., Vol. II, 1979, 479-484 [157] VDI, VDI-GVC: VDI- Wärmeatlas. 8. Aufl., Springer-Verlag, Berlin, 1997 Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

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


131

A

Heat transfer coefficient

A.1 A.1.1

Single phase Inside tube: Turbulent flow Ã a

N u = CRe P r where

Nu = Pr = Re de A P u µw k Cp

hde k Cp µ k ρud µ 4A P

b

µ µw

!c

,

(A.1)

Nusselt number Prandtl number Reynolds number hydraulic diameter cross-sectional area wetted perimeter fluid velocity fluid viscosity at the tube wall temperature fluid thermal conductivity fluid specific heat   

C= 

0.021 gases 0.023 non-viscous liquid 0.027 viscous liquid a = 0.8 b = 0.3 for cooling b = 0.4 for heating c = 0.14

A.1.2

Inside tube: Laminar flow Ã

d N u = 1.86 ReP r L

!1/3 Ã

µ µw

!0.14

,

(A.2)

A.1.3 Shell side For the shell side heat transfer coefficient there are a number of methods the include: • Kern’s method • Donohue’s method • Bell-Delaware method • Tinker’s method Besides these methods there is some proprietary methods putout by various organization for use by their member companies. A number of these method are based on one of the above methods. Some are based upon a judicious combination of methods 3 and 4 above and supplemented by further research data. Among the most popular of the proprietary methods, judged by their large clientele are Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

132

A Heat transfer coefficient

• Heat Transfer Research Inc. (HTRI), Alliambra, california. This method is also known as stream analysis method. • Heat Transfer and Fluid Flow Service (HTFS), Engineering Science Division, AERE, Harwell, United Kingdom Method. In this work only Kern’s method is given below. Bell-Delaware method may be found in Coulson and Richardson’s Ã

N u = 0.36Re where Nu = Pr = Re = de = A= P = G= As = pt = Ds = lB =

hde k Cp µ k Gde µ 4A P

M As (pt −do )Ds lB pt

0.55

Pr

1/3

µ µw

!0.14

,

(A.3)

Nusselt number Prandtl number Reynolds number hydraulic diameter cross-sectional flow area wetted perimeter Mass flux fluid viscosity at the tube wall temperature pitch diameter shell diameter Baffle spacing

Hydraulic diameter (Fig. A.1)    

p2t −πd2o /4 πdo

for square pitch



0.87p2t /2−πd2o /8 πdo /2

for equilateral triangular pitch

de =  

do

pt

pt

As

Square pitch

Equilateral triangular pitch

Cross-flow area

Figure A.1. Tube arrangement


A.2 Condensation

A.1.4

133

Plate heat exchanger Ã

N u = 0.26Re

0.65

Pr

0.4

µ µw

!0.14

,

(A.4)

where Nu = Pr = Re = de = A= P = G= Af = up = M=

A.2 A.2.1

hde = k Cp µ = k ρup de e = Gd µ µ

M Af

=

Nusselt number Prandtl number = Reynolds number hydraulic diameter, taken as twice the gap between the plates cross-sectional flow area wetted perimeter Mass flux cross-sectional area for flow channel velocity mass flow rate

Condensation Condensation on vertical plate or outside vertical tube Ã

k 3 ρ∆ρgλ hm = 0.943 µ∆T L

!1/4

,

(A.5)

where hm = L= k ρ= µ= λ= ∆T = Ts − Tw g= Ts = Tw = A.2.2

mean heat transfer coefficient lenth of the plate or the vetical tube thermal conductivity of the saturated liquid film liquid density liquid viscosity latent heat of evaporization temperature difference across the condensate film acceleration due to gravity saturation temperature of the condensate film wall temperature

Condensation on external horizontal tube Ã

k 3 ρ∆ρgλ hm = 0.725 µ∆T do where

A.2.3

do =

!1/4

,

(A.6)

out side diamter of the tube

Condensation on banks of horizontal tube Ã

k 3 ρ∆ρgλ hm = 0.725 µ∆T Jdo

!1/4

,

(A.7)


134


where

J=

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

In the above equation the condensate film properties save the latent heat of vaporization are evaluated at the film temperature. Tf =

Ts + Tw , 2

(A.8)

the latent heat of vaporization is evaluated at the condensate temperature. For the case of subcooling or superheating the heat transfer coefficient is corrected by substituting the corrected latent heat the heat transfer equation (Rohsenow et al. [121] and Carey [18]) in Nusselt [109]) λ∗ = λ + 0.68cp ∆T . (A.9) A.2.4

Condensation inside horizontal tube Ã

k 3 ρ∆ρgλ hm = 0.555 µ∆T d

A.3

!1/4

,

(A.10)

Two phase flow: Pure fluid

A.3.1 Steiner [140] correlation Steiner [140] has considered the two phase heat transfer coefficient h as a combination of the convective and the nucleate part using an asymptotic model as: ³

h = h3n + 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 types of flow patterns save stratified and stratified-wavy flow) is calculated as 

Ã

 hc ρL = (1 − x) ˙ + 1.2x˙ 0.4 (1 − x) ˙ 0.01  hL0 ρG 



h  G0 x˙ 0.01 1 + 8(1 − x) ˙ 0.7 hL0

Ã

!0.37  +

 !0.67 −2 −0.5  ρL  .  ρG 

(A.12)

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

(ξ/8)(Re − 1000)P r , 1 + 12.7(ξ/8)0.5 (P r2/3 − 1)

(A.13)

taken in to account the respective dimensionless group N uL0 , N uG0 , ReL0 , ReG0 , P rl and P rg . These dimensionless groups are defined as N uL0/G0 =

hL0/G0 d , kL/G

(A.14)


A.3 Two phase flow: Pure fluid

135 md ˙ , µL/G µL/G cp,L/G = . kL/G

ReL0/G0 = P rL/G The friction factor is

(A.15) (A.16)

ξ = (1.82logRe − 1.62)−2 .

(A.17)

For a partial wetting of the tube (stratified or stratified-wavy flow) the average heat transfer coefficient at the tube circumference under the thermal boundary condition of a constant wall temperature is given as hc = hwet (1 − Φ) + hG Φ ,

(A.18)

where hwet is the convective boiling heat transfer coefficient at the wetted part of the tube and it is calculated by using equation A.12. In the non-wetted part of the tube, the convective heat transfer coefficient hg is calculated from the Gnielinski [46] model (equation A.13). In this case Re and N u are defined with the hydraulic diameter of the vapor-occupied part of the tube cross-section Ã

ϕ − sin ϕ dh = d d + 2 sin(ϕ/2)

!

,

(A.19)

where ϕ is the stratified angle. The Reynolds number is given as ReG = and hG =

m ˙ xd ˙ hyd , ²µG

(A.20)

N uG kG . dhyd

(A.21)

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

Ã

!

1/4 x˙ x˙ 1 − x˙ 1.18(1 − x)[gσ(ρ ˙ L − ρG )] ε= (1 + 0.12(1 − x)) ˙ + + 1/2 ρG ρG ρL mρ ˙ .L

)−1

(A.22)

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

fG . fG + fL

(A.23)

With some mathematical manipulation of equation A.23 the non-wetted perimeter can calculated iteratively from the following relationship ϕ = 2πε + sinϕ ,

(A.24)

with the assumption that no bubbles in the liquid phase and no entrainment (hold-up) in the vapor phase, the scaling parameter Φ of equation A.18 can thus be calculated as Φ=

ϕG , 2π

(A.25)

where ϕG = 0.5ϕ. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

136


UG fG

ϕ Ui

d

h

fL UL

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

The local nucleate boiling heat transfer coefficient hnb of a horizontal tube is estimated as Ã !n(pr) hnb q˙ = ψCf F (pr )F (Ra )F (d)F (m, ˙ x) ˙ . (A.26) ho q˙o The value with a subscript ”o” is a reference value. The pressure function is given as Ã

F (pr ) =

2.692p0.43 r

1.6p6.5 r + 1 − p4.4 r

!

,

(A.27)

!0.3  x˙  ,

(A.28)

and the mass flux function is given as 

Ã

q˙ m ˙ 0.25  F (m, ˙ x) ˙ = 1 − p0.1 r m ˙o qcr,nb where The critical value of q˙cr,0,1

q˙cr,cb = 2.79q˙cr,0,1 p0.4 r (1 − pr ) . at a reduced pressure pr of 0.1 is given as

0.25 q˙cr,0.1 = 0.13∆hV,0 ρ0.5 . G,0 [σo g(ρL,0 − ρG,0 )]

(A.29)

(A.30)

The function for the effect of surface roughness and tube diameter is F (Ra ) =(Ra /Rao )0.133 and F (d)=(do /d)0.5 respectively. The pressure dependence of the heat flux exponent n(pr ) can be predicted as n(pr ) = 0.9 − 0.3p0.3 (A.31) r . The experimental value of the specific constant Cf for a number of substances is be found in VDI-Wärmeatlas[157], for example for water it is 0.72. In absence of an experimental value it can be estimated as Ã

f M Cf = 0.789 f MH2

!0.11

,

(A.32)

f is the molecular weight and M f = 2.016. The correction factor ψ for a stratified where M H2 and a stratified-wavy flow pattern under the thermal boundary condition of a constant



137

wall temperature is 0.86 for all other type of flow patterns it is taken as unity (VDIWärmeatlas[157]). Table A.1 shows the reference factors for the nucleate boiling heat transfer coefficient for R134a and R290. Table A.1. Values of the reference parameters used in evaluation of the local nucleate boiling heat transfer coefficient.

Refrigerant

R134a R290

ho W/m2 K

q˙o W/m2

Rao m

do m

3,500 4,000

20,000 20,000

10−6 10−6

0.01 0.01

A.3.2 Kattan et al. [77] correlation For a stratified-wavy flow pattern or annular flow pattern with a partial dryout the two phase heat transfer coefficient is h=

ϕdry hG + (2π − ϕdry )hwet . 2π

(A.33)

The vapor heat transfer coefficient hG is determined by using the Dittus-Boelter [33] correlation as 0.8 0.4 kG hG = 0.023ReG PrG , (A.34) d with Reynold number given as m ˙ xd ˙ ReG = , (A.35) εµG where ε is the void fraction given by the Rauhani [117] model (equation A.22). The heat transfer coefficient on the wetted portion of the tube is q

hwet =

3

h3n + h3c .

(A.36)

The nucleate boiling heat transfer coefficient hn is given by the Cooper [27] model as −0.55 f −05 hn = 55p0.12 M q˙ . r (−0.4343 ln pr )

(A.37)

The convective heat transfer coefficient is given by a modified form of the Dittus-Boelter [33] model as 0.4 kL hc = 0.0133ReL0.69 PrL . (A.38) d 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)


138


where ϕdry is ϕdry = ϕstrat

(m ˙ wavy − m) ˙ , (m ˙ wavy − m ˙ strat )

(A.41)

where ϕstrat is calculated iteratively from equation A.24. The mass flux under a stratified and wavy flow pattern is m ˙ strat =

(226.3)2 fL fG2 ρG (ρL − ρG )µL g cos Θ , 0.3164(1 − x) ˙ 1.75 π 2 µ0.25 L

(A.42)

and (

m ˙ wavy =

"

µ

16fG3 gdρL ρG π2 We F1 (q) ˙ (1 − x) × 2 x˙ 2 π 2 (1 − (2hL − 1)2 )0.5 25hL Fr

¶F2 (q) ˙ L

1 + cos Θ

#)0.5

+ 50 ,

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

F1 (q) ˙ = 646.0

!2

q˙ q˙crit

Ã

+ 64.8

q˙

!

q˙crit

Ã

; F2 (q) ˙ = 18.8

!

q˙ q˙crit

+ 1.023 .

(A.44)

The stratified-wavy flow model is also valid for the stratified flow patten with ϕstrat replacing ϕdry and for the annular flow condition with ϕdry is set to zero and the film thickness δ is set to (1 − ε)d/4. A.3.3 Kandlikar [70] correlation The flow boiling heat transfer coefficient for a pure fluid is given by Kandlikar [70] as h = max(hn , hc ) ,

(A.45)

wher the subscript n and c in equation A.45 refers to the nucleate and convective boiling respectively. The convective and the nucleate boiling part is given as hn = 0.6683Co−0.2 (1 − x) ˙ 0.8 hL0 f (FrL0 ) + 1058.0Bo0.7 (1 − x) ˙ 0.8 FF l hL0 ,

(A.46)

hc = 1.136Co−0.9 (1 − x) ˙ 0.8 hL0 f (FrL0 ) + 667.2Bo0.7 (1 − x) ˙ 0.8 FF l hL0 ,

(A.47)

and

respectively, where F rL0 is the liquid Froude number, Bo is the boiling number and Co is the convection number. These dimensionless groups are defined as F rL0

m ˙ , = ρL gd

q˙ Bo = , m∆h ˙ v

Ã

Co =

ρG ρL

!0.5 µ

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 FF l is a fluid-surface parameter related to the nucleation characteristic. For all type of fluids flowing in a stainless tube it is taken as 1. The single phase heat transfer Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]


139

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 N uL0 =

ReL0 P rL (ξ/2) hL0 d = . 2/3 k 1.07 + 12.7(PrL − 1)(ξ/2)0.5

(A.49)

The Gnielinski [46] correlation (equation A.13) is valid in the range of 0.5 ≤ PrL ≤ 2000 and 2300 ≤ ReL0 ≤ 5 × 104 . The friction factor ξ in equation A.49 is given by equation A.17. A.3.4 Chen [19] correlation Chen [19] postulated that the heat transfer coefficient is made of two parts: a) a microconvective (or nucleate boiling) portion hn and b) a macro-convective (or forced convective) portion hc as h = hc F + hn S , (A.50) where hc is calculated using the Dittus and Boelter [33] correlation as hc = 0.023 where ReL =

kL 0.8 0.4 ReL P rL , d

(A.51)

cpL µL (1 − x) ˙ md ˙ , P rL = , µL kL

(A.52)

The suppression factor for the convection part is

F =

   

1

if 1/Xtt > 0.1

h i0.736    2.35 1 + 0.213 Xtt

, if 1/Xtt ≤ 0.1

and the Martinelli parameter Xtt is given as µ

1 − x˙ X= x˙

¶0.875 Ã

ρG ρL

!0.5 Ã

µL µG

!0.125

.

(A.53)

The nucleate boiling heat transfer coefficient is hn = 0.00122

0.49 kL0.79 c0.45 p,L ρL 0.24 ∆Tsat ∆p0.75 sat , 0.24 0.24 σ 0.5 µ0.29 ρ ∆h L G V

(A.54)

where ∆Tsat = Tw − Ts ;

∆psat = p(Tw ) − p(Ts );

Retp = ReL F 1.25 .

(A.55)

The suppression factor for the nucleate part is S=

1 . 1 + 2.53 × 10−6 Retp

(A.56)


140


A.3.5 Gungor and Winterton [52] correlation The Gungor and Winterton [52] correlation is a modified form of the Chen [19] correlation given by equation A.50 with the nucleate boiling calculated from the Cooper [27] correlation given by equation A.37. The suppression factor for the convection part is defined as  (0.1−2F rL ) 1.16  + 1.37(1/xtt )0.86 )F rL if F r < 0.05  (1 + 24, 000Bo F = ,   1 + 24, 000Bo1.16 + 1.37(1/xtt )0.86 if F r ≥ 0.05 and the suppression factor for the nucleate part is  1/2 2 −1   (1 + 0.00000115F ReL ) F rL

S= 

2

(1 + 0.00000115F ReL )

−1

if F r < 0.05 , if F r ≥ 0.05

The convective boiling part is calculated from the Dittus-Boelter [33] correlation (equation A.51). A.3.6 Shah [130] correlation The Shah [130] correlation is given as h = max(hc , hn ) ,

(A.57)

where the subscript n and c in equation A.57 refers to the nucleate and convective boiling respectively. The convective heat transfer coefficient is defined as hc = 1.8hL N −0.8 , where N=

  

Co

(A.58)

F rL > 0.04 ,

 

0.38F rL−0.4 Co F rL < 0.04

where hL is calculated using the Dittus-Boelter [33] correlation (equation A.51). The nucleate boiling heat transfer coefficient is calculated as follows • For N > 1

  

hn =   • For 1 > N > 0.1 • For N < 0.1

230hL Bo0.5

Bo > 0.0003 .

1 + 46hL Bo

0.5

Bo < 0.0003

hn = F hL Bo0.5 exp(2.74N −0.1 ) .

(A.59)

hn = F hL Bo0.5 exp(2.47N −0.15 ) ,

(A.60)

where F =

   14.7  

Bo > 0.0011 .

15.43 Bo < 0.0011



141

A.3.7 Schrock and Grossman [129] correlation A very simple correlation is given by Schrock and Grossman [129] as "

h = 1.91hL

µ

1 104 × Bo + 1.5 Xtt

¶2/3 #0.6

,

(A.61)

where hL is calculated using Dittus-Boelter [33] correlation equation A.51. A.3.8 Dembi et al. [30] correlation The Dembi et al. [30] correlation is based on the asymptotic model given by equation A.11 with the nucleate and convection part given as kL hn = 23388.5 d

Ã

q˙ ρG ∆hV $

!0.64 Ã

gd ∆hV

and

!0.27 Ã

Ã

i0.11 m ˙ 2 ∆hV kL h 4 hc = 0.115 x˙ (1 − x) ˙ 2 d ρL gσ

m ˙ 2d ρL ∆hV $

!0.14

,

(A.62)

!0.14 0.27 PrL ,

(A.63)

respectively. The parameter $ is defined as $ = 0.36 × 10−3 p−1.4 . r

(A.64)

A.3.9 Klimenko [84] correlation The Klimenko [84] correlation is based on the asymptotic model given by equation A.11 with the convection part given by the Dittus-Boelter [33] correlation equation A.51 and the nucleate boiling is ( hn1 NCB < 1.6 × 104 hn = , hn2 NCB > 1.6 × 104 where

Ã −3

hn1 = 7.4 × 10

hn2 Ã

Pe =

wm b , Rem = νL

kL = 0.087 b

Ã

kw kL

kw kL

!0.15 −1/3

P e0.6 Kp0.5 P rL Ã

!0.09

Re0.6 m

!

ρG ρL

"

Ã

!#

(A.65)

P rL ,

(A.66)

!0.2

qb p , Kp = q , b= ∆hV ρG aL σg(ρL − ρG )

m ˙ ρL wm = 1+x −1 ρL ρG

,

1/6

s

2σ , g(ρL − ρG )

qb , Re∗ = , ∆hV ρG νL

NCB

Rem = Re∗

(A.67)

Ã

!

ρL . ρG (A.68)


142


A.3.10 Jung et al. [64] correlation The Jung et al. [64] correlation is a modified form of the Chen [19] correlation. The convection heat transfer coefficient is calculated using the Dittus-Boelter [33] correlation (equation A.51) and the nucleate part is calculated from the Stephan and Abdelsalm in VDI-Wärmeatlas [157] correlation as "

kL q(b.d) ˙ hn = 207 b.d kL Ts where

#0.745 Ã

ρG ρL

!0.581

Ã

0.533 PrL ,

2σ (b.d) = 0.511 g(ρL − ρG ) µ

1 F = 2.37 0.29 + Xtt   

S= 

A.4

4048Xtt1.22 Bo1.13

(A.69)

!0.5

,

(A.70)

¶

,

(A.71)

Xtt < 1 .

2.0 − 0.1Xtt−0.28 Bo−0.33 1 ≤ Xtt ≤ 5

Two phase flow: Mixture

A.4.1 Steiner [140] correlation Steiner [140] has extended his pure component asymptotic model to mixture. The nucleate part of the heat transfer coefficient is suppressed using the Schl¨ under [126] suppression factor for the nucleate boiling. The Schl¨ under [126] suppression factor is based on the heat and mass transfer laws it is defined as (

"

hid,n Bo q Fn = 1 + (Tb,k − Tb,j )(yej − xej ) 1 − exp q˙ ρL ∆hV βL

#)

,

(A.72)

where Tb is the saturated (boiling) temperature of the pure component, the index j and k stands for the more volatile and less volatile component respectively. βL /B0 = 5 × 105 is the mass transfer coefficient. The ideal nucleate boiling heat transfer coefficient for a mixture hid,n is calculated from the heat transfer coefficient of pure components as "

hid,n =

X x ei

#−1

,

hi,n

(A.73)

and Bo /βL = 5.103 and ρL and ∆hV is the ideal density and enthalpy of evaporation of the mixture respectively. xe and ye is the liquid and vapor mole fraction of the more volatile component respectively. The same approach applies also to the convective part for the liquid-liquid immiscible mixture. That is to say for a liquid-liquid miscible mixture the convective suppression factor made analogous to that for the nucleate boiling heat transfer coefficient as (

"

Bo q hid,c (Tb,k − Tb,j )(yej − xej ) 1 − exp Fc = 1 + q˙ ρL ∆hV βL

#)

.

(A.74)


A.4 Two phase flow: Mixture

143

A.4.2 Kandlikar [71] correlation Kandlikar [71] has extended his pure component correlation (Kandlikar [70]) to mixtures as • Region I: Near-azeotropic region h = max(hn , hc ) ,

(A.75)

where hn and hc is obtained from equation A.77 and equation A.47 respectively using the mixture properties. • Region II: Moderate diffusion-induced suppression region h = hc ,

(A.76)

where hc is given by equation A.77 with the properties of the mixture. • Region III: Severe diffusion-induced suppression region: 0.03< V1 < 0.2 and Bo ≤ 1E −4 ; V1 ≥ 0.2 h = 1.136Co−0.9 (1 − x) ˙ 0.8 hL0 f (FrL0 ) + 667.2Bo0.7 (1 − x) ˙ 0.8 FF l hL0 FD , where

"µ

V1 =

cpL ∆hV

¶µ

a D12

FD =

¶0.5

Ã

dT |ye − xe| dxe

0.678 . 1 + V1

(A.77)

!#

,

(A.78)

(A.79)

A.4.3 Bennett and Chen [8] correlation Bennett and Chen [8] has extended the Chen [19] correlation (equation A.50) for mixture. Here both the convective and the nucleate parts are suppressed. The convection part which is calculated for the original Chen [19] correlation with mixture properties is suppressed using the following suppression factor Fc =

Tw − Tph , Tw − Ts

(A.80)

where Tw , Tph , and Ts is the wall, equilibrium temperature and saturation temperature respectively. The nucleate part is also calculated using the original Chen [19] model for the pure substance with mixture properties. It suppressed using the the suppression factor given by equation A.79. A.4.4 Palen [111] correlation Palen [111] has extended the original Chen [19] correlation for pure component (equation A.50) to mixture similar to the Bennett and Chen [8] correlation. However, only the nucleate part is suppressed using the following suppression factor Fd = exp(−0.027∆Tbp ) ,

(A.81)

where ∆Tbp is difference between the dew and bubble point temperature of the mixture. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

144


A.4.5 Jung et al. [64] correlation Jung et al. [64] have extended their pure substance correlation to the mixture. The nucleate boiling heat transfer coefficient is replaced by the ideal one given by equation A.73. The convective part is suppressed using the following suppression factor Fc = 1.0 − 0 − 35|ye1 − xe1 |1.56 .

(A.82)

For the nucleate part the following suppression factor is employed Fn = where

1 , {[1 + (b2 + b3 )(1 + b4 )](1 + b5 )}2 Ã

1.01 − xe1 b2 = (1 − xe1 ) ln 1.01 − ye1    

!

Ã

xe1 + xe1 ln ye1

0

!

+ |ye1 − xe1 |1.5 ,

Ã

p b4 = 152 pc,1 b5 = 0.92|ye1 − xe1 |

,

!0.66

, Ã

e x1 =1 e y1

(A.84)

x1 ≥ 0.01

b3 =  ³ ´ 0.1  e x1  − 1 xe1 < 0.01 e y1

and

(A.83)

0.001

p pc,1

(A.85) !0.66

,

(A.86)

for xe1 = ye1 = 0 ,

xe1 and ye1 is the liquid and vapor mole fraction of the more volatile component respectively. p and pc,1 is system pressure and critical pressure of the more volatile component respectively.


145

B B.1

Pressure drop Single phase

The pressure drop due to friction exists because of the shear stress between the fluid and the tube wall. Estimation of the friction pressure drop is somewhat more complex and various approaches have been taken, for example the frictional pressure gradient is given as Ã ! 4τo 4f m ˙2 dp = = , (B.1) − dz f d 2dρ where m ˙ is the mass flux in kg/m2 s and f is the friction factor calculated using a Blasiustype model as  0.3164   Re0.25 Re ≥ 2320 f=   64 Re < 2320 . Re Integration of equation B.1 yields ∆p =

B.2

4f m ˙2L , 2ρ d

(B.2)

Two phase

In flow boiling, the temperature drops in the direction of flow as a result of the pressure drop. This results in a change in the driving force (temperature difference) for the heat transfer along the flow path. Thus beside the heat transfer coefficient, knowledge of the pressure drop is of paramount importance in the design of the evaporator. In the present work the pressure drop is measured simultaneously with the heat transfer coefficient along the test section. The momentum balance implies that the two phase pressure gradient is composed of three components as Ã ! Ã ! Ã ! dp dp dp dp = + + , (B.3) dz dz f dz a dz h where dp/dz, (dp/dz)f , (dp/dz)a and (dp/dz)h is the total, friction, acceleration and hydrostatic pressure gradient respectively. For a horizontal tube the hydrostatic pressure gradient diminishes. The acceleration pressure drop is caused by the change in momentum in both the liquid and vapor phases. The change in the momentum stems from the change in the velocity of the two phases, which is brought about by the added (or withdrawn) heat to/from the test section. For the case of adiabatic flow the acceleration pressure drop diminishes for ∆pa /ps → 0 (Baehr and Stephan [3]), where ps is the saturation pressure. There exist in the literature a number of approaches for modelling the change in the static pressure drop due to acceleration. The most widely accepted models include homogenous or separated flow models. The separated flow model is also widely known as the heterogenous model. In the homogenous model the static pressure drop due to acceleration is " Ã ! # Ã ! 1 1 1 dp 2 d =m ˙ x˙ − + . (B.4) − dz a dz ρL ρG ρL The energy balance in a small unit length dz along the test tube yields 4q˙ dx˙ = . dz m∆h ˙ vd

(B.5)


146

B Pressure drop

Substitution of equation B.5 into equation B.4 yields the pressure drop due to acceleration as Ã ! 4q˙m ˙ ρG ∆pa = 1− ∆L . (B.6) d∆hv ρG ρL In the separated flow model the static pressure drop due to acceleration can be derived from the momentum balance as Ã

dp − dz

!

"

d x˙ 2 (1 − x) ˙ 2 =m ˙ + dz ερG (1 − ε)ρL 2

a

#

.

(B.7)

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

−∆pa = −(po − pi )a = m ˙

2

(1 − x˙ o )2 x˙ 2i (1 − x˙ i )2 x˙ 22 + − − εo ρG,o (1 − εo )ρL,o εi ρG,i (1 − εi )ρL,i

#

.

(B.8)

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

Ã

!

1/4 x˙ x˙ 1 − x˙ 1.18(1 − x)[gσ(ρ ˙ L − ρG )] ε= (1 + 0.12(1 − x)) ˙ + + 1/2 ρG ρG ρL mρ ˙ L

)−1

,

(B.9)

where ρL and ρG is the liquid and vapor density respectively, which are calculated from the fundamental equation of state of Tillner-Roth and Baehr [152] for R134a. g is acceleration due to gravity, σ is the surface tension, m ˙ is the mass flux and x˙ is the quality. The surface tension is calculated using the method of Lucus [92] given in VDI-Wärmeatlas [157]. The pressure drop due to friction exists because of the shear stress between the fluid and the tube wall. Estimation of the friction pressure drop is somewhat more complex and various approaches have been taken, for example in homogenous or separated flow models. In the homogenous model the frictional pressure gradient is given as Ã

dp − dz

!

= f

2ξ m ˙2 4τo = , d dρH

(B.10)

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

 0.3164   Re0.25  

64 Re

Re

≥ 2320

Re < 2320 .

and the homogenous densityρH is given as 1 − x˙ x˙ 1 = + . ρH ρL ρG

(B.11)

The two phase Reynolds number Re is Re =

md ˙ , ηT P

(B.12)

where ηT P is a two-phase viscosity. A variety of methods have been proposed to calculate the two phase viscosity, a commonly used one being that proposed by McAdams et al. [95] 1 ηT P

=

1 − x˙ x˙ + , ηL ηG

(B.13)


B.2 Two phase

147

where ηL and ηG are the liquid and vapor viscosity. In the separated flow model the two phase frictional pressure drop is related to that for single phase as Ã Ã ! ! dP dp = ΨG/L , (B.14) dz f dz f,L/G where Ψ is the two phase multiplier. There exist a number of correlations for the prediction of Ψ. These include Friedel [42], Chishlom [22] and Lockhart and the Martinelli [91] model. These models are presented in Appendix B. There exists a number of correlations for the prediction of the two phase multiplier Ψ of the separated flow model. These models are presented in the following subsections. B.2.1

Friedel [42] model ΨL0 = E +

3.24F H F r0.045 W e0.035

,

(B.15)

where E = (1 − x) ˙ 2 + x˙ 2

ρL fG0 , ρG fL0

(B.16)

F = x˙ 0.78 (1 − x) ˙ 0.24 , Ã

H=

ρL ρG

!0.91 Ã

µG µL

Fr =

!0.19 Ã

µG 1− µL

(B.17) !0.7

,

m ˙2 , gdρ2H

(B.18) (B.19)

m ˙ 2d , (B.20) σρH d is tube diameter, σ is the surface tension and %H is the homogenous density given by equation B.11. fG0 and fL0 are the friction factors defined by a Blasius-type model as We =

fL0/G0 =

0.079 , Re0.25 L0/G0

(B.21)

where Re = md/µ. ˙ The range of the validity of the Friedel [42] model is µL /µG < 1000 B.2.2 Lockhart and Martinelli [91] model In the Lockhart and Martinelli [91] model the two phase friction multiplier is C 1 + 2 , X X

(B.22)

2 ψG = 1 + C.X + X 2 ,

(B.23)

ψL2 = 1 +

where X is the Martinelli parameter and the value of the coefficient C is given in Table B.1. The range of the applicability of the Lockhart and Martinelli [91] correlation is µL /µG >1000 and m ˙ <100 kg/m2 s. Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

148

B Pressure drop

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

Liquid

Gas

Subscript

C

Turbulent Viscous Turbulent Viscous

Turbulent Turbulent Viscous Viscous

tt vt tv vv

20 12 10 05

B.2.3 Chisholm [22] model In the Chisholm [22] model the two phase friction multiplier is h

i

ΨL0 = 1 + (Y 2 − 1) B x˙ (2/n−1) (1 − x) ˙ (2/n−1) + x˙ 1−n , where Y2 =

(dpf /dz)G0 , (dpf /dz)L0

(B.24)

(B.25)

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

55 m ˙ 1/2

0 < Y < 9.5 ,

(B.26)

520 9.5 < Y < 28 , (B.27) Ym ˙ 1/2 15000 B = 2 1/2 28 < Y . (B.28) Y m ˙ The range of the validity of the Chisholm [22] correlation is µL /µG > 1000 and m ˙ > 100 kg/m2 s. B=


149

C

Physical properties

The fluid physical properties required for heat exchanger design are divided in thermodynamic and trasport properties. The transport properties include viscosity, thermal conductivity, surface tension and diffusion coefficient are generally calculated from the existing correlations (Pery and Coulson). The thermodynamic properties include demsity, specific heat temperature, pressure (vapor), enthalpy, latent heat of evaporation. Beside the fluid properties the thermal conductivity of the material is necessary for the evaluation of heat transfer coefficient. The thermodynamic properties are evaluated using critical tables.

C.1

Physical properties: Pure fluid

C.1.1 Specific heat The specific heat of the ideal gas is given in as Cp = CP V AP A + (CP V AP B)T + (CP V AP C)T 2 + (CP V AP D)T 3

(C.1)

Where T is in K and CPVAPA, CPVAPB, CPVAPC, CPVAPD are constant in ideal gas heat capacity. These constant are given in Appendix A for organic and inorganic compounds. C.1.2 Vapor pressure The vapor pressure is generally predicted using Antonie equation as ln p = AN T A −

AN T B T + AN T C

(C.2)

where T is in K and ANTA, ANTB,ANTC are Anonie equation constant. These constant are given in Appendix D for organic and inorganic compounds. C.1.3 Liquid viscosity The liquid viscosity is given as: µ

1 1 log µ = V ISA − T V ISB

¶

(C.3)

where VISA, VISB are constants in the liquid viscosity equation. These constant are given in Appendix D for organic and inorganic compounds. C.1.4 Vapor dynamic viscosity VDI-W¨ armeatlas [157] Lucas and Luckas [92] in VDI-Wärmeatlas [157] have recommended the following procedure for the calculation of the vapor viscosity. η = (ηξ)r Fp FQ

1 , ξ

(C.4)

for Tr ≤ 1 and pr ≤ ps /pc

with

(ηξ)r = 0.600 + 0.760pαr + (6.990pβr − 0.6)(1 − Tr ) ,

(C.5)

and β = 1.390 + 5.746pr , α = 3.262 + 14.98p5.508 r

(C.6)


150

C Physical properties

for 1≤ Tr ≤ 40 and 0≤ pr ≤ 100

"

ApE r (ηξ) = (η ξ) 1 + −1 BpFr + (1 + CpD r ) r

o

#

,

(C.7)

where η o is the low pressure viscosity given as η o ξ = [0.807Tr0.618 − 0.357 exp(−0.449Tr ) + 0.340 exp(−4.058Tr ) + 0.018]Fpo FQo , (C.8) and ξ is given as [Tc ]1/6 [R]1/6 [Na ]1/3 ξ= , (C.9) [M ]1/2 [pc ]2/3 where Na is the Avagadro number in kmol. The coefficients of equation C.7 are given as a1 A = exp(a2 Trγ ) , (C.10) Tr B = A(b1 Tr − b2 ) , (C.11) c1 C = exp(c2 Trδ ) , (C.12) Tr d1 D = exp(d2 Tr² ) , (C.13) Tr E = 1.3088 , (C.14) ς F = f1 exp(f2 Tr ) . (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 b1 f1

1.245.10−3 1.6553 0.9425

a2 b2 f2

5.1726 1.2723 −0.1853

c1 d1 ς

Fp = 1 +

(Fpo

0.4489 c2 1.7368 d2 0.4489 ² "

and

"

(ηξ)r − 1) ηoξ #−1

3.0578 2.2310 -7.6351

γ δ

-0.3286 -37.7332

#−3

, "

Ã

(C.16) !#4

(ηξ)r (ηξ)r FQ = 1 + − 1) − 0.007 ln , (C.17) ηoξ ηoξ where Fpo and FQo is low-pressure polarity and quantum factors respectively. These factors are Fpo = 1 , 0 ≤ µr < 0.022 , (C.18) (FQo

Fpo = 1 + 30.55(0.292 − Zc )1.7 , 0.022 ≤ µr < 0.075 ,

(C.19)

Fpo = 1 + 30.55(0.292 − Zc )1.7 (|0.96 + 0.1(Tr − 0.7)|) , 0.075 ≤ µr , (C.20) o where Zc is the critical compressibility factor and FQ = 1.0 for all substances other than He, H2 and D2 . The reduced dipole moment µr is given as µr =

µ2 pc , (kTc )2

(C.21)

where the dipole moment µ for the gases is given in VDI-Wärmeatlas [157] Dr. Ali A. Rabah, Dept of Chemeng, U of K, Email : [email protected]

C.1 Physical properties: Pure fluid

151

C.1.5 Dynamic viscosity of Fenghour et al. [40] The functional form of the liquid and vapor viscosity of ammonia as given by Fenghour et al. [40] is η = ηo (T ) + η1 (T )ρ + η2 (ρ, T ) , (C.22) The first term of the expansion is the dilute gas term which is given as ·

¸

f T )1/2 0.021357 (M ηo (T ) = 100 , 0.29572 exp(Ω)

(C.23)

f is the molecular weight in g/mol, T is the temperature in K. The collision where M integral Ω is defined as (

Ã

kT Ω(T ) = C(1) + C(2) log ²

!

"

4 X

Ã

kT + C(n) log ² n=3

!#n )

,

(C.24)

where ²/k=386 K and the value of the coefficient C is given in table C.2. Table C.2. Coefficients for the Collision integral Ω (equation C.24).

C(1)

4.9931822

C(2) -0.61122364

C(3) 0.18535124

C(4) -0.1116094

The second term of equation C.22 represents the contribution of the moderately dense fluid η1 (T ) = Fv (T )ηo (T )ρ , where

  

13 X

 

i=2

Fv (T ) = C A(1) +

"

Ã

A(i) log

(C.25)

 !# −(i−1)   2 kT

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 3 5 7 9 11 13

-0.17999496×101 -0.53460794×103 -0.13019164×105 -0.58711743×105 -0.59834012×105 -0.12027350×105 -0.120807957×103

2 4 6 8 10 12

0.466692621×102 0.33604074×104 0.33414230×105 0.71426686×105 0.33652741×105 0.24348205×104

The third term in the viscosity equation C.22 is the contribution of the dense gas η2 (ρ, T ) =

3 X

F (i, T )ρi+1 ,

(C.27)

i=1


152


where

 ³ ´2 ³ ´4 ² ² −1   1 0.219664285 − 0.83651107 × 10  kT kT       ³ ´   ² −2 −2   2 0.17366936 × 10 − 0.83651107 × 10  kT 

F (i, T ) = 

³ ´2 ³ ´3   ² ² −3 −3  3 0.167668649 × 10 − 0.149710093 × 10 +   kT kT       ³ ´4    0.77012274 × 10−4 kT²

The Fenghour et al. [40] correlation for the vapor viscosity of ammonia has an uncertainty of 2% in the temperature range of T < Tc . C.1.6 Surface tension Lucas and Luckas [92] in VDI-Wärmeatlas [157] have recommended the following correlation for the calculation of the surface tension µ ¶m 2/3 1/3 1 − Tr σ = pc Tc b, (C.28) a where the reduced pressure and temperature are defined as pr =

p T , Tr = , , pc Tc

(C.29)

respectively. For a polar fluid like R134a the following quantities are valid a b m X

= = = =

1, 0.1574 + 0.359ω − 1.769X − 13.69X 2 − 0.510ω 2 + 1.298ωX , 1.210 + 0.5385ω − 14.61X − 32.07X 2 − 1.656ω 2 + 22, 03ωX , lgpsr (Tr = 0.6) + 1.70ω + 1.552 .

(C.30) (C.31) (C.32) (C.33)

where ω is the acentric factor and it is given by Pitzer in VDI-Wärmeatlas [157] as The surface tension given by equation C.28 is in 10−5 N/cm. Its level of uncertainty as given by Reid et al. [118] is 1.2 % in the range of the reduced temperature of 0.56 ≤ Tr ≤ 0.63. C.1.7

Thermal conductivity for liquids µ

¶

ρ 1/3 k = 3.65 × 10 Cp . (C.34) M where k thermal conductivity W/moC, M is the molecular mass, Cp speific heat capacity 3 (kJ/kg oC), ρ density (kg/m ) −5

C.1.8

Thermal conductivity for gases µ

¶

10.4 . (C.35) M where k thermal conductivity W/m o C, M is the molecular mass, Cp specific heat capacity (kJ/kg o C), µ viscosity in (mNs/m2 ) k = µ Cp +


C.2 Physical properties: Mixture

153

C.1.9 Specific enthalpy For the vapor phase, the deviation of the specific enthalpy from the ideal state can be illustrated using Redlich-Kwong equation written as z 3 + z 2 + z(B 2 + B − A) = 0 .

(C.36)

where z is the compressibilty factor defined as pv . RT

z= and A=

aP

h = ho + RT +

C.2

,B =

R2 T 2.5

(C.37) bp . RT

Z v "Ã 0

dP T 2 2.5 R T dT

(C.38)

!

#

− p dv .

(C.39)

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 X

xi . ln ηL,i + 2.xe1 .xe2 .G12 ,

(C.40)

i=1

where xei is the mole fraction of the component i, ηL,i is the viscosity of the component i in kg/ms and G12 is an adjustable parameter normally obtained from experimental data. For a polar-nonpolar mixture G12 = -0.22. The Reid et al. [118] model give the thermal conductivity with a mean error of less then 5%. C.2.2 Vapor dynamic viscosity of mixtures The viscosity of a gas mixture can be approximated by using the principle of the kinetic theory (Reid et al. [118]) as o ηm = ηm + ∆η , (C.41) o where ηm is the mixture gas viscosity at a low pressure and ∆η is a correction factor for the high pressure viscosity n X yei ηG,i o , (C.42) ηm = Pn ei φij j=1 y i=1

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

φij =

f /M f )0.25 1 + (ηG,i /ηG,j )0.5 (M j i f /M f )]0.5 [8(1 + M i j

φji =

f ηG,j M j φ . f ij ηG,i M i

i2

,

(C.43)

(C.44)


154


The high pressure correction term is estimated as i

h

∆η =

0.497.10−6 exp(1.439ρr,m ) − exp(−1.111ρ1.858 r,m ) 1/6

−2/3

f −0.5 pc,m Tc,m M m

.

The pseudo critical properties of the mixture are calculated as X X X pc,j υc,j Tc,m = yej Tc,j , υc,m = yej υc,j , Zec,j = , Zem = yj Zec,j , RT c,j j=1 j j f = M m

X

f, yej M j

ρc,m =

j=1

f /1000 M m , υc,m

ρr,m =

ρm , υc,m

pc,m =

(C.45)

(C.46)

RTc,m Zec,m , (C.47) υc,m

where T is in K, p is in Mpa, υc,m is in m3 /kmol, ρr,m is in kg/m3 , M is in g/mol and ηm is in kg/ms. The error associated with this model is seldom exceeded 3 to 4% (Perry and Green [112]). C.2.3 Liquid thermal conductivity of mixtures Reid et al. [118] have recommended a Filippov-like model for the prediction of the thermal conductivity of a liquid mixture as λm =

2 X

f λ − 0.72X X f X i L,i 1 2 |λL,2 − λL,1 | ,

(C.48)

i=1

f and X f is the weight fraction of the component 1 and 2 respectively and λ and where X 1 2 1 λ2 is the thermal conductivity of the component 1 and 2 in W/mK respectively.

C.2.4 Vapor thermal conductivity of mixtures The thermal conductivity of a low-pressure gas mixture can be determined from the relationship given by Reid et al. [118] λG,m =

n X

yei λG,i , Pn ei Aij j=1 y i=1

(C.49)

where λG,m is the low-pressure gas mixture thermal conductivity, λG,i is the low-pressure thermal conductivity of the pure component i. For a binary mixture of two non-polar gases or a non-polar and a polar gas, Aij may be calculated by the model given by Perry and Green [112] as h i c /M f )0.25 2 1 + (λtr,i /λtr,j )0.5 (M j i Aij = , (C.50) f /M f )]0.5 [8(1 + M i j with Γj exp(0.0464Tr,i ) − exp(−0.2412Tr,i ) λtr,i = , (C.51) λtr,j Γi exp(0.0464Tr,j ) − exp(−0.2412Tr,j ) f is the molecular weight and Γ is defined as where M "

f3 Tc,i M i Γi = 210 Pci4

#(1/6)

,

(C.52)

f is in g/mol and λ is in W/mK. This model yields an error where T is in K, p is in bar, M of less than 5% in the prediction of the thermal conductivity of the gas mixture.


C.3 Software packages

155

C.2.5 Surface tension of mixtures Lucas and Luckas [92] in VDI-Wärmeatlas [157] recommended the following method for calculation of the mixture surface tension µ

σm = where

2/3 1/3 pc,m Tc,m

1 − tr,m am

¶nm

bm ,

(C.53)

#

"

X Ts,ri ln(pc,m /1.01325) , bm = xei bi , bi = 0.1196. 1 + 1 − Ts,ri

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

X

xei Tc,j , υc,m =

X

j=1

Zem =

X j

xj υc,j , Zec,j =

j

xej Zec,j ,

pc,m =

RTc,m Zec,m , υc,m

(C.54)

pc,j υc,j , RTc,j

Ts,ri =

(C.55)

Tb,i , (C.56) Tc,i

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

C.3

Software packages

There exists a number of software packages for the prediction of thermodynamic and transport properties. These include: 1. ASPEN Plus ( 2. CHEMCAD 3. SUPERPRO 4. REFPROP


course_Heat exchanger_04

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