The Concise Valve Handbook, Volume II: Actuation, Maintenance, and Safety Relief [1 ed.] 9781947083691, 9781947083684

Table of contents :
Cover
Contents
List of Figures
List of Tables
Foreword
Chapter 7: Valve Actuators and Positioners
Chapter 8: Valve Testing and Diagnostics
Chapter 9: Valve Maintenance and Repair
Chapter 10: Safety Relief Valves
Appendix A: J–T Valve
Appendix B: Basic Acoustics
Appendix C: Block and Bleed
Appendix D: Water Hammer
Appendix E: Stainless Steel
Glossary
Bibliography
About the Author
Index
Adpage
Backcover

Citation preview

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The Concise Valve Handbook Actuation, Maintenance, and Safety Relief, Volume II Michael A. Crabtree

CRABTREE

EBOOKS FOR THE ENGINEERING LIBRARY

AUTOMATION AND CONTROL COLLECTION

Research studies within the process industry routinely indicate that the fluid control valve is responsible for 60 to 70% of poorfunctioning control systems. Furthermore, valves in general are consistently wrongly selected, regularly misapplied, and often incorrectly installed. This two-volume book comprises a comprehensive up-to-date body of knowledge that provides a total in-depth insight into valve

THE CONTENT

and actuator technology—looking not just at control valves, but a

• Manufacturing Engineering • Mechanical & Chemical Engineering • Materials Science & Engineering • Civil & Environmental Engineering • Advanced Energy Technologies

whole host of other types including: check valves, shut-off valves, solenoid valves, and pressure relief valves.

Whilst studying the correct procedures for sizing, readers will also learn the correct procedures for calculating the spring ‘wind-up’ or ‘bench set’. Maintenance issues also include: testing for deadband/ hysteresis, stick-slip and non-linearity; on-line diagnostics; and signature analysis. Written in a detailed but understandable language, the two volumes are presented in a form suitable for both the beginner, with no prior knowledge of the subject, and the more advanced

THE TERMS • Perpetual access for a one time fee • No subscriptions or access fees • Unlimited concurrent usage • Downloadable PDFs • Free MARC records For further information, a free trial, or to order, contact:  [email protected]

specialist. For the last sixteen years, ‘Mick’ Crabtree, who holds an MSc in industrial flow measurement, has been involved in technical training and consultancy—running workshops on industrial instrumentation and networking throughout the world covering the fields of process control (loop tuning), process instrumentation, data communications,

The Concise Valve Handbook, Volume II

A methodology is presented to ensure the optimum selection of size, choice of body and trim materials, components, and ancillaries.

The Concise Valve Handbook Actuation, Maintenance, and Safety Relief Volume II

fieldbus, safety instrumentation systems (according to both ISA S84 and IEC 61508/61511), project management, on-line analysis, and technical writing and communications. This book represents some thirty years wealth of experiential knowledge gleaned by the author working within a wide variety of industries and from more than 6000 technicians and engineers who have attended the author’s workshops. ISBN: 978-1-94708-369-1

Michael A. Crabtree

The Concise Valve Handbook

The Concise Valve Handbook Actuation, Maintenance, and Safety Relief Volume II

Michael A. Crabtree

MOMENTUM PRESS, LLC, NEW YORK

The Concise Valve Handbook: Actuation, Maintenance, and Safety Relief, Volume II Copyright © Momentum Press®, LLC, 2018. All rights reserved. No part of this publication may be reproduced, stored in a retrieval system, or transmitted in any form or by any means—­ electronic, mechanical, photocopy, recording, or any other—except for brief quotations, not to exceed 400 words, without the prior permission of the publisher. First published by Momentum Press®, LLC 222 East 46th Street, New York, NY 10017 www.momentumpress.net ISBN-13: 978-1-94708-369-1 (print) ISBN-13: 978-1-94708-368-4 (e-book) Momentum Press Automation and Control Collection Cover and interior design by Exeter Premedia Services Private Ltd., Chennai, India 10 9 8 7 6 5 4 3 2 1 Printed in the United States of America

To my wife Pam—for her love and patience

Abstract Volume II: Actuation, Maintenance, and Safety Relief, takes an in-depth look at actuators and positioners. This volume also explores a variety of maintenance and diagnostic issues including: testing for dead-band/­ hysteresis, stick-slip and non-linearity; on-line diagnostics; signature analysis; and correct procedures for calculating the spring “wind-up” or “bench set.” A complete section is also devoted to the whole field of safety relief devices. Lastly, this volume covers a number of topics which are all too often ignored: acoustics; water hammer; and even classification of stainless steel.

Keywords acoustics, actuators, block and bleed, diagnostics, fail-safe, maintenance, positioners, safety relief valves, stainless steel classification, transfer mechanisms, water hammer

Contents List of Figures

xv

List of Tables

xxiii

Foreword

xxv

Volume I 1   Basic Principles

1

1.1  The Final Control Element as Part of the Control Loop

2

1.2  Basic Theory

3

1.3  Equation of Continuity

3

1.4  Bernoulli’s Equation

5

1.5  Choked Flow

8

1.6  Pressure Recovery

9

1.7  Turndown Ratio and Rangeability

11

1.8  Velocity Profiles

12

1.9  Reynolds Number

13

1.10 Flashing and Cavitation

14

1.11 Flashing

15

1.12 Cavitation

16

1.13 Leakage Classification

18

1.14 Isolation Valve Leakage Classification

21

2   Liquid Valve Sizing

23

2.1  Practical Considerations

23

2.2  Application of Formulae

24

2.3  Sizing Example 1

27

x  •  Contents

2.4  Piping Geometry Factor

29

2.5  Sizing Example 2

31

3   Gas Valve Sizing

33

3.1  Pressure Drop Mechanism

33

3.2  Specific Heat Ratio Factor

38

3.3  Gas Expansion Factor

40

3.4  Valve Sizing

41

3.5  Sizing Example 1

43

4   Valve Construction

47

4.1  Globe Valve

48

4.2  Bonnet Assembly

49

4.3  PTFE (Teflon)

49

4.4  Laminated Graphite

50

4.5  Extended Bonnet

52

4.6  Bellows Seal Bonnet

52

4.7  Valve Trim

54

4.8  Guiding

55

4.9  Post-guiding

55

4.10 Top- and Bottom-guided Double Seat

56

4.11 Single-ported Balanced Globe Valve

57

4.12 Cage-guiding

58

4.13 Split Body Globe

59

4.14 Angle Is

60

4.15 Needle Valve

61

4.16 Bar Stock Body Valve

61

4.17 Gate Valve

61

4.18 Wedge Gate

62

4.19 Slab Valve

64

4.20 Expanding Gate Valve

65

4.21 Knife Edge Gate Valve

67

4.22 Pinch Valve

69

4.23 Diaphragm Valve

72

4.24 Rotary Control Valves

74

Contents   •   xi

4.25 Ball Valve

75

4.26 Trunnion Ball Valve

77

4.27 Characterized Ball Segment Valve

81

4.28 Butterfly Valve

82

4.29 Plug Valve

84

4.30 Eccentric Plug Valve

86

4.31 Check Valves

88

4.32 Valve Sizes and Pipe Schedules

88

4.33 Material Selection

90

4.34 Corrosion

90

4.35 Erosion

94

4.36 End Connections

94

4.37 Screwed End Connections

94

4.38 Flanged End Connections

95

4.39 Hub End Body

96

4.40 Welded End Connections

96

4.41 Lap Joint Flange

98

4.42 Flangeless Connections

99

4.43 Grayloc® Connector 5   Valve Trim and Characterization

100 103

5.1  Inherent Characteristics

103

5.2  Linear Inherent Flow Characteristic

103

5.3  Equal Percentage Inherent Flow Characteristic

104

5.4  Quick Opening Inherent Flow Characteristic

104

5.5  Modified Percentage Inherent Flow Characteristic

105

5.6  Characteristic Profiling

105

5.7  Installed Characteristics

105

5.8  Cavitation Control

108

5.9  Reducing Cavitation

110

5.10 Eliminating Cavitation

112

5.11 Noise Sources

113

5.12 Mechanical Noise

115

5.13 Hydrodynamic Noise

116

xii  •  Contents

5.14 Aerodynamic Noise

117

5.15 Noise Prediction

117

5.16 Noise control

117

5.17 Path Treatment

118

5.18 Insulation

119

5.19 Silencers

120

5.20 Source Treatment

120

5.21 Velocity Control

120

6   Valve Selection

123

Glossary

129

Bibliography

131

About the Author

133

Index

135

Volume II 7    Valve Actuators and Positioners 7.1  Pneumatic Control 7.2  Flapper–Nozzle Assembly 7.3  I/P Converter 7.4  Diaphragm Actuators 7.5  Springless Diaphragm 7.6  Advantages and Disadvantages of Diaphragm Actuators 7.7  Cylinder Actuators 7.8  Spool Block 7.9  Electro-Hydraulic Actuation 7.10 Electric Actuation 7.11 Torque Limiting 7.12 Hammer-Blow Mechanism 7.13 Solenoid Valve 7.14 Digital Actuators 7.15 Transfer Mechanisms 7.16 Valve Positioners 7.17 Positioner Guidelines

141 141 141 142 144 145 147 147 149 149 150 152 153 153 155 157 161 163

Contents   •   xiii

8   Valve Testing and Diagnostics

167

  8.1  Deadband and Hysteresis

167

  8.2  Testing Procedures

169

 8.3  Online Diagnostics

173

 8.4  Electronic Torque Monitoring

176

9   Valve Maintenance and Repair

179

 9.1   In-Line Repairs

180

 9.2   Repairs Under Pressure

180

  9.3   Repairs on Drained Systems

181

 9.4   Packing Replacement

181

  9.5   Replacing or Refinishing Seat Rings

181

 9.6   Other In-Line Repairs

182

 9.7   In-Line Post-Repair Procedures

182

 9.8   Shop Repairs

182

 9.9   Actuator Bench Set

183

 9.10  Spring Calculations

184

10  Safety Relief Valves

187

  10.1  History

187

  10.2  Definitions

190

  10.3  Weight-Loaded Pressure/Vacuum Relief Valves

191

 10.4  Spring-Loaded Relief Valves

192

 10.5  Applications

194

 10.6  Limitations

194

 10.7  Safety Valves

194

  10.8  Basic Operation: Lifting

196

  10.9  Basic Operation: Reseating

198

  10.10 Conventional Safety Relief Valves

200

  10.11 Balanced Safety Relief Valves

204

  10.12 Bellows-Type Balanced Safety Valve

204

  10.13 Piston-Type Balanced Safety Valve

206

  10.14 Non-Reclosing Pressure Relief Devices

211

  10.15 Conventional Rupture Disc

215

  10.16 Scored Tension-Loaded Rupture Disc

215

xiv  •  Contents

  10.17 Composite Rupture Disc

216

  10.18 Graphite Rupture Disc

217

  10.19 Burst Disc Applications and Installation Practices

218

  10.20 Performance Tolerance

219

  10.21 Maximum Operating Pressure

221

  10.22 Cyclic/Pulsating Duties

221

  10.23 Case A: 276 kPa (g) or Higher

222

  10.24 Case B: Lower than 276 kPa (g)

223

  10.25 Standards

223

Appendix A: J–T Valve

225

Appendix B: Basic Acoustics

227

Appendix C: Block and Bleed

241

Appendix D: Water Hammer

245

Appendix E: Stainless Steel

255

Glossary

263

Bibliography

265

About the Author

267

Index

269

List of Figures Figure 7.1. The flapper–nozzle assembly converts a small physical displacement into a pressure change.

142

Figure 7.2. As the flapper moves away from the nozzle, the airflow  will increase and the output pressure will fall. Although this produces a non-linear output, over the normal range of interest, 0.2 to 1 bar (3 to 15 psi), the relationship is normally considered to be a straight line. 143 Figure 7.3. The 4–20 mA current signal is applied to a coil that is physically attached to a spring diaphragm on which is mounted the flapper. As the current varies, its magnetic field interacts with the permanent magnet field, ­deflecting the diaphragm by an amount proportional to the control signal current. 143 Figure 7.4. Typical configuration of an I/P converter (courtesy Foxboro).

144

Figure 7.5. A direct-acting diaphragm actuator.

145

Figure 7.6. A reverse-acting diaphragm actuator.

146

Figure 7.7. Springless diaphragm actuator uses a differential air input.

146

Figure 7.8. The cylinder or piston-type actuator.

148

Figure 7.9. Size comparison between a diaphragm actuator (left) and cylinder actuator (right) mounted on two comparable valves (courtesy Valtek International).

148

Figure 7.10. Pneumatic spool assembly in which the incoming air is supplied to either one or other side of the actuator, while at the same time, simultaneously exhausting air from the opposite side (courtesy Mitech). 149 Figure 7.11. Swing jet controller.

150

xvi  •   List of Figures

Figure 7.12. Basic electrically operated actuator.

151

Figure 7.13. (a) The worm drive is free to move longitudinally on a spline that is held in its central position by means of ­pre-loaded torque springs. (b) If the valve reaches its end position, the tangential force on the driven wheel rises and displaces the worm gear axially on its shaft against the pressure of the holding springs.

153

Figure 7.14. The worm wheel and output shaft are connected via a dog coupling with backlash. When the rotation direction is reversed, the backlash first has to be covered, and the motor can run up to its nominal output speed without load before the valve is unseated (hammer blow).

154

Figure 7.15. A direct-acting solenoid valve.

154

Figure 7.16. A three-way solenoid valve might be used to switch air from one side of an actuator diaphragm to the other 155 (a) de-energized position (b) energized position. Figure 7.17. P&ID representation of a three-way solenoid valve (dot indicates normally open (N.O.) port and a solid ­indicates normally closed (N.C.) port.

155

Figure 7.18. Typical construction of a ‘four-phase’ stepping motor with a basic step of 1.8°.

156

Figure 7.19. Rack and pinion transfer mechanism (courtesy Mitech).157 Figure 7.20. Double-piston arrangement (a) air is supplied f­ orcing the pistons away from each other and rotating the ­pinion ­anticlockwise (b) air exhaustion (loss of pressure) allows compressed springs to force the pistons toward each other and rotate the pinion clockwise (courtesy Spirax Sarco). 158 Figure 7.21. Double crank transfer mechanism (courtesy Mitech).

159

Figure 7.22. In the double-crank mechanism, the run torque (the torque in a mid-position) is higher than the end torque.

159

Figure 7.23. Scotch yoke transfer mechanism (courtesy Mitech).

160

Figure 7.24. Vector diagrams showing reaction force on the rocker arm.

161

Figure 7.25. Vector diagrams showing moment of the reaction force.161

List of Figures   •   xvii

Figure 7.26. In the scotch yoke mechanism, the end torques are twice as high as the run torque (the torque in a mid-position).161 Figure 7.27. Basic principle of operation of a pneumatic positioner. 162 Figure 8.1. Deadband as a result of mechanical play within a gear-train.

168

Figure 8.2. An illustration of hysteretic error.

168

Figure 8.3. Hysteresis: A combination of deadband and hysteretic error.

168

Figure 8.4. Determining the effects of hysteresis/deadband as a result of an input change of two steps up, three down and one up (courtesy Michael Brown Control Engineering).169 Figure 8.5. The effects of stick-slip, without hysteresis and deadband (courtesy Michael Brown Control Engineering).170 Figure 8.6. As the PD increases in regular step, the PV increases in a series of steps that gradually become smaller, showing a marked non-linearity that is typical of an oversized valve.

171

Figure 8.7. Testing connections for a complete pneumatically operated final control assembly.

172

Figure 8.8. Negative hysteresis: one of the effects of an undersized actuator (courtesy Michael Brown Control Engineering).

173

Figure 8.9. Plotting the ‘valve signature’ plot with the actuator pressure plotted on the y-axis and the travel plotted along the x-axis. The separation between the opening (red) and closing (blue) lines is the result of the friction band (courtesy Fisher-Emerson).

174

Figure 8.10. The packing friction is approximately twice that of the previous example. Typically, this might be due to errantly over-tightening the packing, resulting in the ­excessive friction (courtesy Fisher–Emerson).

175

Figure 8.11. An example of a valve signature showing several revealed faults (courtesy Fisher–Emerson).

176

Figure 8.12. Opening torque characteristics of a typical wedge gate valve in which the valve position (travel) is plotted on the x-axis and the torque demand is plotted on the y-axis (courtesy Rotork). 177

xviii  •   List of Figures

Figure 9.1. Although not generally recommended, in-line repair can be carried out while the line is still under pressure, on gate and globe valves having back seats. 180 Figure 9.2. Forces developed on a nominal 50-mm valve plug.

185

Figure 9.3. Unbalance of forces.

186

Figure 10.1. Family of pressure relief devices, classified as either reclosing or non-reclosing.

188

Figure 10.2. Papin’s safety valve was kept closed by means of a lever and movable weight. Sliding the weight along the lever kept the valve in place and regulated the steam pressure.

188

Figure 10.3. The weight of the seat assembly (or pallet) keeps the valve closed until the pressure acting on the underside equals this weight.

191

Figure 10.4. When associated with tank breathing (1 to 4 in WC), weight-loaded valves are often referred to as a ‘conservation valves’ and provide IN- and OUT-breathing.192 Figure 10.5. Spring-loaded pressure relief valve.

193

Figure 10.6. The valves have closed bonnets to prevent the release of corrosive, toxic, flammable, or expensive fluids and can be supplied with lifting levers. 194 Figure 10.7. The spring of a safety valve is usually fully exposed and has a lifting lever for manual opening.

195

Figure 10.8. Illustration of the standard defined areas.

196

Figure 10.9. Typical disc and shroud arrangement used on rapid opening safety valves.

197

Figure 10.10. Operation of a conventional safety valve.

197

Figure 10.11. Relationship between pressure and lift for a typical safety valve.

198

Figure 10.12. Blowdown ring is threaded around the valve nozzle and positioned to form a huddling chamber with the disc skirt.

199

Figure 10.13. When the blowdown ring is adjusted up, the forces required to lift the seat disc occur very close to set pressure and the blowdown is long. When the ring is adjusted down, the seat lift does not occur until the ­pressure under the seat disc is considerably higher and the blowdown is short. 199

List of Figures   •   xix

Figure 10.14. Schematic diagram of a valve with the spring housing vented to the discharge side of the valve. 201 Figure 10.15. Schematic diagram of a valve with spring housing vented to the atmosphere.

202

Figure 10.16. Bellows-type balanced safety valve.

205

Figure 10.17. Block schematic of a bellows-type balanced safety valve showing force balancing.

205

Figure 10.18. Block schematic of piston-type balanced safety valve showing force balancing. 207 Figure 10.19. High-pressure pilot-operated valve incorporating an unbalanced piston and an integrally mounted pilot.

208

Figure 10.20. Alternative seating arrangements available for a pilot-operated piston-type safety relief valve.

209

Figure 10.21. Low-pressure diaphragm-type pilot-operated valve.

209

Figure 10.22. Similar in construction to a spring-loaded valve, but using a shear-pin in place of a spring.

212

Figure 10.23. The basic buckling pin valve comprises a pin of a precise length that holds a piston on its seat. As the pressure increases and the axial force on the pin subsequently also increases, the pin will buckle.

212

Figure 10.24. Also known as a rupture disc, bursting disc, or burst diaphragm, the disc is designed to rupture at a ­predetermined pressure and, once ruptured, will not re-seal (courtesy Oseco).

214

Figure 10.25. Bursting discs may be forward- or reverse-acting.

215

Figure 10.26. Typical rupture disc holders (courtesy Oseco).

215

Figure 10.27. Conventional domed rupture discs are pre-bulged solid metal discs designed to burst when operating conditions are 70% or less of the rated burst pressure.216 Figure 10.28. Scored tension-loaded rupture discs allow a closer ratio (generally 85%) of system operating pressure to disc burst pressure.

216

Figure 10.29. Composite rupture disc (courtesy Continental Disc Corp.).

217

Figure 10.30. Graphite rupture disc manufactured from graphite impregnated with a binder material and designed to burst by bending or shearing (courtesy Svi Carbon Private Limited).

217

xx  •   List of Figures

Figure 10.31. Bursting disc installed on a safety valve.

219

Figure 10.32. Performance tolerances.

220

Figure 10.33. A zero-manufacturing range is the tightest, meaning the average of the burst tests in the factory must equal the nominal burst pressure at the coincident temperature.

221

Figure 10.34. Determining the upper and lower maximum operating ranges according to whether the pressure lies above or below 276 kPa (g). 223 Figure B.1. Graphic representation of a sound wave showing frequency and amplitude.

228

Figure B.2. Addition of harmonics to the fundamental: (a) second harmonic; (b) third harmonic.

229

Figure B.3 Two identical loudspeakers connected across an amplifier.

231

Figure B.4. Logarithmic response of the human ear.

232

Figure B.5. The 10-fold power ratio increase is designated a Bel with each power increment of 26% being one-tenth of a Bel—called a decibel (dB).

233

Figure B.6. Simple circuit showing power developed by a resistor.

235

Figure B.7. Doubling the power is achieved by only a 1.414 increase in voltage.

236

Figure B.8. Threshold of hearing (lower) and of pain (upper).

238

Figure B.9. Equal loudness contours.

239

Figure B.10. A- B-, and C-weighted responses required for measuring sound pressure levels.

240

Figure C.1. (a) Under normal operation, the valves are set with the isolation valves 1 and 2 open and the bleed valve 3 closed. (b) When isolating the downstream equipment, the valves are set with isolation valves 1 and 2 closed and bleed valve 3 open.

242

Figure C.2. Typical construction of a single double block-and-bleed valve (courtesy Habonim).

242

Figure C.3. The bleed often takes the form of a cap or plug.

243

List of Figures   •   xxi

Figure D.1. If a moving column of liquid (a) is slowed down suddenly by, for example, a quick-closing valve, the sudden change in liquid velocity in the delivery line creates a pressure wave (b).

246

Figure D.2. The pressure wave travels back up the line (a) at between 1,000 and 1,300 m/s, to the end of the pipe where it will reverse direction and travel back toward the valve (b).

247

Figure D.3. Depending on the valve size and system conditions, a valve closing in 1.5 s or less can produce a pressure spike five times the system working pressure. 248 Figure D.4. Hydraulic shock wave produced as a result of accumulated condensate in steam piping.

249

Figure D.5. A pulsation dampener or surge suppressor. During a surge, the fluid pressure displaces the bladder and compresses the trapped gas.

251

Figure D.6. A Daniel gas-loaded axial flow style valve in which nitrogen gas is used to pressurize the valve piston to keep it in the closed position (courtesy Emerson).

252

Figure D.7. As the pipeline pressure increases, the combined force of the spring and nitrogen gas pressure is overcome and the valve opens (courtesy Emerson).

253

Figure E.1. In ferritic stainless steel, the atoms are arranged in a body-centered cubic structure.

256

Figure E.2. Ferritic stainless steel family.

257

Figure E.3. With the addition of nickel, the atoms in austenitic stainless steels are arranged on the corners of the cube and also in the center of each of the faces.

257

Figure E.4. The relationship between the various 300 series austenitic grades.

258

Figure E.5. The martensitic Grade 400 series.

261

Figure E.6. Relationship between the complete family of stainless steels.

262

List of Tables Table 10.1. Typical operating pressure to burst pressure ratios (courtesy BS&B) Table 10.2. Safety valve performance summary Table B.1. The wavelengths of several frequencies travelling in air Table B.2. Gain and attenuation ratios expressed in dBs Table B.3. Sound intensities of various sources Table D.1. Some typical velocities in various liquids Table E.1. Difference in the properties of ferritic and austenitic stainless steels Table E.2. Chemical composition of standard grades (courtesy International Stainless Steel Forum)

222 224 230 234 238 246 258 260

Foreword In this book, ‘The Concise Valve Handbook—Part 2. Actuation, Maintenance, and Safety Relief,’ I have made use of a building-block approach, presenting material in a form suitable for two distinct classes of reader: the beginner, with no prior knowledge of the subject and the more advanced specialist. The complete text is suitable for the advanced reader. However, those parts of the text that involve a mathematical treatment, which are not required by the beginner, are indicated by a mark ► at the beginning and ◄ at the end. Consequently, for the beginner, the text may be read, with full understanding, by ignoring the marked sections. I offer no apologies for my preference for metric-based measurement— the SI system. Apart from the United States, only two other countries in the world still adhere to the fps system (foot-pound-second)—the so-called imperial system first defined in the British Weights and Measures Act of 1824—Burma and Liberia. I have tried to mix it up as far as possible, and I have got a units conversion table right in the front of the book. But, for the moment, just try for the following: 1 bar = 100 kPa ≈ 1 atmosphere ≈ 14.7 psi 1 inch = 25.4 mm 20 °C = 68 °F 100 °C = 212 °F And lastly, while I have made some compromises (analog instead of analogue; program instead of programme), I reserve my right to spell according to the British system:

xxvi  •  Foreword

English Metre Litre Fibre Colour

United States Meter Liter Fiber Color

Unit Conversions Quantity Distance

Area

Volume

Mass Force Pressure

SI 25.5 mm 1 millimetre 1m 1m 0.9144 m 1 square metre (m2) 1 square metre (m2) 1 square metre ­millimetre (mm2) 1 cubic metre (m3) 1 cubic metre (m3) 0.02832 m3 1 litre 1 litre 1 litre 3.785 litres 1 kg 454 g 1N 4.448 N 1 bar 1 kPa (kN/m2) 6.895 kPa 1 psi

United States customary 1 in 0.03937 in 39.37 in 3.281 ft 1 yd 1550 in2 10.76 ft2 0.00155 in2 61.02 in3 35.31 ft3 1 ft3 61.02 in3 0.03531 ft3 0.2642 gal 1 gal 2.205 lb 1 lb 0.2248 lbf 1 lbf 14.504 lbf/in2 (psi) 0.145 lbf/m2 (psi) 1 psi 0.0361 inches H2O (in WC)

Foreword   •   xxvii

Quantity Temperature Flow rate

SI K °C 1 m3/h 1 kg/h

United States customary 1.800 °R 1.8 °C + 32 = °F 4.403 gal/min (gpm) 2.205 lb/h

CHAPTER 7

Valve Actuators and Positioners In any flow control loop, a primary sensing flow device is used to produce a signal, which ultimately controls a valve: either to open or close, in an ON/OFF mode, or to provide proportional control. The actuator, therefore, is that part of the final control element that moves the control valve— either in a linear manner (for the control of a globe valve) or in a rotary manner (for control of butterfly or ball valves). An actuator may be powered electrically, pneumatically, or hydraulically. However, despite the trend away from pneumatically controlled instrumentation and toward electronics, the actuator still remains predominantly pneumatic.

7.1 Pneumatic Control In the process control instrumentation field, the pneumatically controlled actuator is still used for four main reasons: • users feel that there is little, if any, improvement in the performance of electric actuators; • the cost of electric actuators is higher than pneumatic actuators; • the power dissipation on electric actuators is considered excessive, giving rise to thermal problems; and • users feel that pneumatic control is more reliable and can provide a FAIL-OPEN or FAIL-CLOSE operation.

7.2 Flapper–Nozzle Assembly At the heart of most pneumatic process control systems lies the flapper– nozzle assembly (Figure 7.1)—a device that converts a small physical

142  •   The Concise Valve Handbook Supply pressure

Pressure reducing restriction

Flapper

Output pressure Nozzle b a

Figure 7.1.  The flapper–nozzle assembly converts a small physical displacement into a pressure change.

displacement into a pressure change. An air supply is applied to the nozzle via a pressure reducing restriction, such that the output pressure will be lower that the supply pressure by an amount determined by the flow of air from the nozzle. The outflow of air from the nozzle varies according to the position of the flapper. With the nozzle covered, the pressure approaches the supply pressure, but as the flapper moves away from the nozzle, the airflow will increase and the output pressure will fall. This is shown in Figure 7.2. The input displacement is applied to the flapper, which increases or decreases the distance from the nozzle, and thus varies the output pressure. It should be noted that the displacement range is quite small, of the order of micrometres and produces a non-linear output. However, over the normal range of interest, 0.2 to 1 bar (3 to 15 psi), the relationship is normally considered to be a straight line.

7.3 I/P Converter A common application of the flapper–nozzle device is the electro-­ pneumatic current-to-pneumatic converter, normally referred to as an I/P converter, which, typically converts a standard 4–20 mA process signal into a pneumatic output varying linearly over the range 20 to 100 kPa (3 to 15 psi). In a practical arrangement, the flapper is physically attached to a spring diaphragm on which is mounted a coil system (Figure 7.3). The coil is arranged within a magnetic field that is produced by a permanent magnet. As the current in the coil varies from 4 to 20 mA, it produces

Valve Actuators and Positioners   •  143 2.0 1.8

Pressure (bar)

1.6 1.4 1.2 1.0 0.8 0.6 0.4 0.2 0

a

Displacement (µm)

b

Figure 7.2.  As the flapper moves away from the nozzle, the airflow will increase and the output pressure will fall. Although this produces a non-linear output, over the normal range of interest, 0.2 to 1 bar (3 to 15 psi), the relationship is normally considered to be a straight line.

Permanent magnet 4 – 20 mA

Coil

Nozzle

Flapper

Spring diaphragm

Figure 7.3.  The 4–20 mA current signal is applied to a coil that is physically attached to a spring diaphragm on which is mounted the flapper. As the current varies, its magnetic field interacts with the permanent magnet field, deflecting the diaphragm by an amount proportional to the control signal current.

a magnetic field that interacts with the permanent magnet field. The ­diaphragm, thus, deflects by an amount proportional to the control signal current, to produce a change in the flapper–nozzle gap. The fully packaged I/P converter (Figure 7.4) comprises the flapper/ nozzle assembly together with a downstream volume booster that acts as a pilot-operated regulation device.

144  •   The Concise Valve Handbook Permanent magnet Coil Spring diaphragm

4 – 20 mA

Flapper Nozzle Exhaust to atmosphere

Exhaust to atmosphere Pilot air

Supply air

Diaphragm Control diaphragm Control air

Figure 7.4.  Typical configuration of an I/P converter (courtesy Foxboro).

The supply air is applied to the lower chamber of the volume booster where a certain amount, determined by the position of the control diaphragm, flows to the output. When the flapper moves closer to the nozzle, the dynamic back-­ pressure increases until it corresponds to the input pressure and pushes both the diaphragm and the control diaphragm downward, causing the output pressure to increase until a new state of equilibrium is reached in the diaphragm chambers. When the output pressure decreases, the diaphragm moves upward, allowing the output pressure to vent until the forces on the diaphragms are balanced again.

7.4 Diaphragm Actuators The diaphragm actuator is the most widely used pneumatic actuator for proportional control. As shown in Figure 7.5, the variable operating air is applied to one side of a flexible diaphragm. In this form, the lower chamber is vented to atmosphere and the operating air, thus, moves the diaphragm downward, against the force of the ‘ranging’ spring. It is important, at this point, to consider the effect of air pressure failure. Since the plug stem needs to move upward to open the valve and increase the flow, the direct-actuating diaphragm would be closing the valve against the range spring pressure—and the pressures acting on the plug of the valve. Thus, in the event of air pressure failure, the valve would go to a fully open position—fail-open.

Valve Actuators and Positioners   •  145 Travel stop

Control air input

Spring

Diaphragm

Spring flange

Actuator stem

OPEN

Travel indicator plate

Figure 7.5.  A direct-acting diaphragm actuator.

In many cases, the fail-open mode is highly desirable. There are, however, also many more cases where a fail-close mode of operation is required. As shown in Figure 7.6, in the reverse-acting diaphragm actuator, the variable operating air is applied to the lower sealed chamber with the upper chamber vented to atmosphere. In this case, failure of the air supply would result in the actuator stem moving downward under the action of the spring.

7.5 Springless Diaphragm In the springless diaphragm actuator (Figure 7.7), the control air is applied differentially to both sides of the diaphragm. This arrangement allows a much higher actuating force to be applied, for emergency on/off control, since one side can be bled, and there is no restraining opposition, other than the valve itself. The springless diaphragm may also be used for proportional control with the signal air pressure fed to one side of the diaphragm and a separate regulated supply fed to the other side.

146  •   The Concise Valve Handbook

Spring

Diaphragm Travel stop

Control air input

Actuator stem

OPEN

Travel indicator plate

Figure 7.6.  A reverse-acting diaphragm actuator.

Air

Air

OPEN

Diaphragm

Figure 7.7.  Springless diaphragm actuator uses a differential air input.

Valve Actuators and Positioners   •  147

7.6 Advantages and Disadvantages of Diaphragm Actuators The main advantage of the diaphragm-type actuator is its cost since it is the least expensive method of applying proportional control. In addition, because it is the most widely used type of actuator, a wide choice of devices is available to suit virtually any type of valve. Furthermore, by using a characterized spring, it is possible to obtain rough control by feeding a 0.2 to > 1 bar (3–15 psi) signal directly onto the diaphragm. Further advantages of the diaphragm actuator are: it is essentially fail-safe; the speed is adjustable from fast to slow, with reasonably close control; it is easily adapted for use in explosion proof areas; and it is easy to maintain. One of the main problems of the diaphragm actuator becomes ­apparent when high-thrust forces are required, for example, to obtain tight shut-offs on certain types of valves. To obtain high thrusts, either the diaphragm area or the control pressure must be increased, both of which place increased restraints on the casing. Thus, for example, the Samson type 271 pneumatic actuator can have an effective diaphragm area of up to 2,800 cm2, but with a maximum pressure of 3 bar to provide a thrust of some 84 kN. Already, however, the outside diameter of the casing is some 600 mm, a bulky and top-heavy structure when used with smaller valves. Other disadvantages include: stiffness is low, and so, precise control is not always possible; the need for a supply of clean air; and, if required, hand-wheel overrides are expensive and large.

7.7 Cylinder Actuators The cylinder or piston-type actuator (Figure 7.8) makes use of a cast cylinder much better able to withstand high pressures than the diaphragm type (up to 1 MPa) and may be hydraulically or pneumatically operated. To generate a thrust of 84 kN (as in the previous example), the piston area must only be 840 cm2—almost half the diameter. Figure 7.9 shows a size comparison between a diaphragm actuator (left) and cylinder actuator (right) mounted on two comparable valves. Although many cylinder actuators are spring-opposed, they are generally used either in a differential mode or make use of a constant load air cushion restraint.

148  •   The Concise Valve Handbook Control input

Piston

Spring Cylinder

Air cushion Yoke

Neoprene boot Travel indicator

Figure 7.8.  The cylinder or piston-type actuator.

Figure 7.9.  Size comparison between a diaphragm actuator (left) and cylinder actuator (right) mounted on two comparable valves (courtesy Valtek International).

Generally, spring-and-diaphragm actuators contribute less friction to the control valve assembly than piston actuators, and their frictional characteristics are more uniform with age. Piston actuator friction will probably increase significantly with use as guide surfaces and the O-rings wear, lubrication fails, and the elastomer

Valve Actuators and Positioners   •  149

degrades. Thus, to ensure continued good performance, maintenance is required more often for piston actuators than for spring-and-diaphragm actuators. If that maintenance is not performed, process variability can suffer dramatically without the operator’s knowledge.

7.8 Spool Block In the springless piston-type actuator, in which the air must be applied differentially, the pneumatic amplifier often takes the form a spool assembly (Figure 7.10). Movement of the spindle switches the incoming air supply to either one or other side of the actuator, while at the same time, simultaneously exhausting air from the opposite side. The spindle, which moves inside the spool block, must be virtually frictionless to ensure that the spindle will move for small changes in the input signal.

7.9 Electro-Hydraulic Actuation The increasing acceptance of electronics in the process control industry has led to the need for direct electronic control of actuators. One such development is the electro-hydraulic actuator used to control, for example, a cylinder actuator. One such operating device, the swing jet, is illustrated in Figure 7.11. High-pressure oil flows through a pivoted needle jet, which may be deflected by coils to one side or the other. In the quiescent Spool block

Spindle

Supply top of actuator

Exhaust top of actuator Supply in

Supply in

Exhaust bottom of actuator

Supply in

Supply bottom of actuator

Figure 7.10.  Pneumatic spool assembly in which the incoming air is supplied to either one or other side of the actuator, while at the same time, simultaneously exhausting air from the opposite side (courtesy Mitech).

150  •   The Concise Valve Handbook High pressure oil input

Pivot Deflection coils Swinging needle jet Pick-up block

A

Outlet

B

Figure 7.11.  Swing jet controller.

state, the high-pressure oil impinges in the center of the pick-up block and the outlet oil pressure from A and B is equal. Using electronic control, the jet may be deflected to the left or right to increase the flow into either of the output pick-ups A or B. Such a differential output would, therefore, result in movement of the piston-type actuator, with position feedback ensuring that the needle is returned to its central quiescent position when the valve stem reaches its demanded position. The main drawback of this design is that, due to the pressure losses within the swinging jet itself, the full power of the hydraulic pressure is not realized at the actuator. Although this problem may be overcome by using a hydraulic servo, this serves to increase the cost further, a cost that is already higher than pneumatically operated diaphragm actuators.

7.10 Electric Actuation One of the major disadvantages of the electro-hydraulic-type system is the need for a constant source of pressure, entailing the constant use of electric power: the constant running of motor and pumps. This problem is overcome in the electrical actuator (often referred to as a Motor-operated

Valve Actuators and Positioners   •  151

(MOV)) where an electric motor drives the valve stem through a worm gear assembly (Figure 7.12). Voltage requirements are generally in the range 110/230 V AC and 24 V AC or 24 V DC. One of the fundamental requirements of any actuator/valve combination is that it should be non-reversing. The motor should drive the valve and the forces on a butterfly valve, for example, must not feed back and drive the motor. Traditionally, this non-reversing characteristic has been accomplished using a simple worm gear system in which a worm drive on the motor shaft drives a worm gear. In this arrangement, the wheel cannot drive back through the worm as long as the worm crosses the wheel at an acute angle (less than 45°). Such gearing also provides a speed reduction of as much as 100:1—with a corresponding increase in torque—providing a reasonably compact solution for even large valves. Despite these advantages, electric actuators suffer from a number of drawbacks that preclude their use in all, but a few applications. One of the most serious limitations of an electric actuator is the speed of valve movement, which can be as low as 4 s/mm, and generally rules out their use for modulating control. Another serious a drawback is that MOVs generally have a ‘fail-inplace,’ rather than a ‘failsafe’ action. For this reason, most are equipped with a mechanically operated hand-wheel that allows the valve to be manually operated to its open or closed position in the event of power

Worm gear drive

Motor

Figure 7.12.  Basic electrically operated actuator.

152  •   The Concise Valve Handbook

or mechanical failure. A number of solutions are also offered in which a spring or battery-backed UPS drives the valve to its open or closed position in the event of power failure neither of which is entirely satisfactory. Apart from being, generally, both cumbersome and expensive, springs require the electric motor to be up to three times the size otherwise required, and batteries do not deliver power quickly enough to provide the rapid shutdown needed on an electric actuator to replicate spring closure. Another solution lies in the use of electric actuators incorporating advanced ‘super’ capacitors that can be easily configured to drive a valve to any position (open, close, or any intermediate position) on loss of power or control signal. A further problem with electric actuators is that there are only a few that are certified for hazardous areas. Lastly, electric actuators are generally more expensive than their equivalent pneumatically operated actuators, and their complexity makes them more difficult to maintain.

7.11 Torque Limiting The basic role of the actuator is to move the valve to either a mechanically limited end position or an intermediate position. At the same time, so as not to overload the valve, the actuator must avoid producing any excess torque either during the travel or at the end positions. Thus, an important consideration in the design in any actuator is to ensure that the torque drive is discontinued when the valve reaches its end limits (fully opened or fully closed). Many designs of electrical actuator accomplish this by limit switches. However, if faced with the possibility of a limit switch failure, precautions must be taken to ensure that, when an end limit is reached and the torque starts to rise, that it does not increase to a point where the valve is damaged. In order not to overload the valve, the actuator must avoid producing any excess torque either during the travel or at the end positions. In one method, the worm drive is free to move longitudinally on a spline and is held in its central position by means of pre-loaded torque springs (Figure 7.13). If now, while the drive is running the valve should reach its end position (or become jammed), then the tangential force on the driven wheel will rise considerably. This rise in torque displaces the worm gear axially on its shaft against the pressure of the holding springs. This movement is detected by means of a lever that operates the torque switch.

Valve Actuators and Positioners   •  153 Splined shaft

Worm drive

Pre-loaded torque springs

Worm gear

(a)

(b)

Figure 7.13.  (a) The worm drive is free to move longitudinally on a spline that is held in its central position by means of pre-loaded torque springs. (b) If the valve reaches its end position, the tangential force on the driven wheel rises and displaces the worm gear axially on its shaft against the pressure of the holding springs.

7.12 Hammer-Blow Mechanism Often, rotary valves that are seldom operated become jammed or sticky and are difficult to open or close. In many cases, this problem may be overcome by the application of a quick, focused blow (similar to that of a hammer striking an anvil). One method of applying such a controlled blow, used by Auma Riester GmbH & Co. is shown in Figure 7.14. The worm wheel and output shaft are connected via a dog coupling with backlash. When the direction of rotation is reversed, the backlash first has to be covered and the motor can run up to its nominal output speed without load before the valve is unseated (hammer blow).

7.13 Solenoid Valve The solenoid valve is essentially an on–off control element comprising an electromagnetically operated plunger (or core) that is attached directly to the valve stem (Figure 7.15). Available in normally open or normally closed configurations, solenoid valves are widely used for emergency shut-off service or for opening a valve simultaneously with the operation of, for example, a pump.

154  •   The Concise Valve Handbook Driven plate Drive plate Drive shaft

Driven shaft

Drive dog attached to drive plate

Driven shaft Driven plate ‘Backlash’ slot in driven plate

Figure 7.14.  The worm wheel and output shaft are connected via a dog coupling with backlash. When the rotation direction is reversed, the backlash first has to be covered, and the motor can run up to its nominal output speed without load before the valve is unseated (hammer blow).

Figure 7.15.  A direct-acting solenoid valve.

When used in safety shut-off applications, the valve is normally held in its energized position and would open or close to allow air to bleed from, or be supplied to, the actuator.

Valve Actuators and Positioners   •  155 Spool

Common

Normally open

Flow Path when de-energized

Return spring

Solenoid

Normally closed Flow Path when de-energized

Normally open

Normally closed

Figure 7.16.  A three-way solenoid valve might be used to switch air from one side of an actuator diaphragm to the other (a) de-energized position (b) energized position.

s

Figure 7.17.  P&ID representation of a three-way solenoid valve (dot indicates normally open (N.O.) port and a solid indicates normally closed (N.C.) port.

A three-way solenoid valve (Figures 7.16 and 7.17) might be used to switch air from one side of an actuator diaphragm to the other.

7.14 Digital Actuators Digital type actuators are centered around the stepping motor, which is, essentially, a DC motor in which the output shaft can be made to move in a series of discrete angular steps. This is achieved through a special motor design combined with an electronic drive circuit that applies

156  •   The Concise Valve Handbook

current pulses to a series of windings in a rotating sequence. The actual methods of construction are many and varied, but the most commonly available motor is known as ‘four-phase’ and has a basic step angle of 1.8° (200 steps/r). This is achieved by using a laminated iron rotor incorporating a permanent magnet that has 50 teeth, which are accurately machined (Figure 7.18). The stator has eight toothed pole pieces, each with its own winding, and the clearance between stator and rotor is extremely small. By means of a Vernier arrangement, the stator windings can be sequentially energized to produce 200 steps in a revolution. Motors are also available with other step angles, for example, 32, 48, 100, and 500 steps per revolution. As an actuator, the stepping motor may be applied in three ways: direct mechanical drive, stepper motor control of a pneumatic or hydraulic servo valve, or stepping motor hybrid positioner in conjunction with a piston actuator. To date, little progress has been made in digital actuators with a sufficiently low price tag, sufficient power, and sufficient speed to capture a large share of the market.

5-tooth pole piece Stator winding 50-tooth rotor lamination stack

Figure 7.18.  Typical construction of a ‘four-phase’ stepping motor with a basic step of 1.8°.

Valve Actuators and Positioners   •  157

7.15 Transfer Mechanisms Most rotary actuators use a linear cylinder in conjunction with a transfer mechanism to translate the linear movement of the piston into rotary motion of the shaft. Three mechanisms are commonly used:

7.15.1 Rack and Pinion The rack and pinion transfer mechanism is effectively a car-steering mechanism in reverse in which (Figure 7.19) a pinion gear is attached to the drive shaft, the rack is attached to the pistons (most of these designs use a double-piston arrangement), and the movement of the pistons causes the shaft to rotate. While this is a compact and neat arrangement, there are several disadvantages: only one tooth is fully engaged; the size is limited; there is no adjustable end stop in the closed position of spring return units; the unit comprises a complex multi-spring pack, and therefore maintenance is not easy; no hand-wheel override is possible; and the mechanism is prone to wear due to continuous reversals, hence backlash.

Figure 7.19.  Rack and pinion transfer mechanism (courtesy Mitech).

158  •   The Concise Valve Handbook Pistons

Pinion driven anticlockwise

(a)

Air in

Air out

Pinion driven clockwise

(b)

Figure 7.20.  Double-piston arrangement (a) air is supplied forcing the pistons away from each other and rotating the pinion anticlockwise (b) air exhaustion (loss of pressure) allows compressed springs to force the pistons toward each other and rotate the pinion clockwise (courtesy Spirax Sarco).

The disadvantage of single-gear engagement is overcome in the double-piston arrangement (Figure 7.20). However, these are even more complex. 7.15.2 Double Crank Figure 7.21 shows how a rocker plate is attached rigidly to the drive shaft and an arm connects the piston shaft to the rocker plate. This arm rotates to take up the lateral movement of the rocker plate pivot joint as the shaft rotates. Major advantages are that the run torque (i.e., the torque in a mid-­ position) is higher than the end torque (Figure 7.22).

Valve Actuators and Positioners   •  159

Figure 7.21.  Double crank transfer mechanism (courtesy Mitech). 800

Torque (N.m)

700 600 500 400 300 200 100 10

20

30

40

50

60

70

80

90

Shaft angle (°)

Figure 7.22.  In the double-crank mechanism, the run torque ( the torque in a mid-position) is higher than the end torque.

This is an advantage for modulating control applications and for large valves on liquid applications where dynamic torques are significant. In addition, backlash is negligible due to the fact that all movement takes place at pivot points in bearings, unlike other designs that incorporate gears or sliding surfaces. This is an advantage for modulating control duties. However, it must be recognized that this is a more expensive design due to the larger number of components, and that in most applications, the actuator is sized on the end torque that is produced by the unit. For this

160  •   The Concise Valve Handbook

reason, it is often necessary to use a larger model for the double-crank design than for the scotch yoke. 7.15.3 Scotch Yoke With this design (Figure 7.23), a pin is assembled through the piston shaft and a slot is machined in the rocker arm to take up the relative lateral movement of the rocker arm and the piston shaft as the arm rotates. While this design does not appear to differ significantly from the ­double-crank mechanism, the characteristic is surprisingly different. The reason is that, in the end positions (Figure 7.24), the pin acts like a wedge in the slot, which in those positions is at approximately 45° to the piston shaft. Since there can be no force on the rocker arm in the direction of the slot, the force produced by the piston must be opposed by a force perpendicular to the slot. The vertical component of the reaction must be the same as the piston force, and so, the reaction force must equal F/cos 45 when the rocker arm is at 45° to the piston shaft. Note that there must also be a reaction in the bearings equal to F. If the horizontal distance between the center line of the piston shaft and the center line of the drive shaft is M, then the moment arm that the above force works on is M/cos 45° (Figure 7.25). Thus, the torque produced at the end position is: (F/cos 45°) . (M/cos 45°) = 2 FM This means that, in theory, this type of actuator transfer mechanism produces twice the torque at the end (Figure 7.26) than it does in the center position. The high-end torque characteristic of the scotch yoke is ideal for on–off duties for ball and butterfly valves where the greatest torque is required to seat and unseat the ball or disc. This usually results in a smaller unit than any of the other mechanisms. In addition, large actuators of this design are possible. A major disadvantage is that the run torques are low compared with the end torque, and so, this type of actuator is not suitable for moduFigure 7.23.  Scotch lating control. In addition, more wear can be yoke transfer mechexpected on the piston shaft bushes due to the anism (courtesy Mitech). high side thrusts.

Valve Actuators and Positioners   •  161 Piston center line Pin

Piston F

at d en

os

45

°)

Effective length in center (M)

Figure 7.24.  Vector diagrams showing reaction force on the rocker arm.

Torque (N.m)

/c

h

Pivot

(M

gt

45°

er

en

is

nt

el

F/cos 45°

ax

ce

iv

ot

op

ct

R

Sl

Drive shaft center line

fe

ea

St

F

Ef

ns 45°

o cti

Pin

Pivot

Figure 7.25.  Vector diagrams showing moment of the reaction force.

600 500 400 300 200 100 10

20

30

40

50

60

70

80

90

Shaft angle (°)

Figure 7.26.  In the scotch yoke mechanism, the end torques are twice as high as the run torque (the torque in a mid-position).

7.16 Valve Positioners In any proportional control system, it is assumed that a given signal output represents a corresponding control action. In the case of an actuator, a given signal air input should result in a specific opening of the valve. Thus in the linear range 20–100 kPa, 60 kPa represents a 50% control action, which should be reflected in the valve moving to a 50% open position. However, the valve may not assume its correct position for a number of reasons. These may include valve stem friction, unbalanced forces on valve plug, variations in the process fluid pressure, etc. As shown earlier, I/P transducers generally provide operating pressure in the 20–100 kPa range, and thus 100 kPa is the maximum force that could be applied to an actuator diaphragm. By contrast, the valve

162  •   The Concise Valve Handbook

positioner acts as a pneumatic relay that is capable of applying the full force of the supply air (from 550 up to as much as 700 kPa) to drive the valve, and thus overcome the various forces that prevent the valve from reaching its correct position. Figure 7.27 illustrates the basic principle of operation. The signal control air is fed to the proportioning bellows that creates a signal force on one end of a pivoted lever: the flapper. The resulting change in the flapper–nozzle gap will raise or lower the control pressure applied to the diaphragm of the air relay. This action positions the pilot valve to either feed or vent the supply air to the actuator, which moves the valve stem up or down. This action, in turn, is fed back via the feedback positioner to balance the force at the opposite end of the flapper lever from the proportioning bellows. Many devices include a cam that can be used to characterize the system and compensate for non-linearity in the process and/or final control devices. It should be noted that, although positioner cams can be used to modify the characteristic of the final control element, its effect is limited in most cases. This is because the cams can also dramatically change the positioner loop gain, which severely limits its dynamic response. However, while characterizing the valve trim is far more effective, the use of cams is always better than no characterization at all and is often the only choice with rotary valves. Many modern electronic positioners incorporate valve characterization that electronically shapes the input signal ahead of the positioner

Control diaphragm Control input

Actuator diaphragm

Proportioning bellows Nozzle

Flapper

Restriction

Supply

Feedback cam

Figure 7.27.  Basic principle of operation of a pneumatic positioner.

Valve Actuators and Positioners   •  163

loop. Such devices use a pre-programmed table of values to produce the valve input required to achieve the desired valve characteristic. This is sometimes referred to as ‘forward path’ or ‘set-point characterization.’ Because characterization occurs outside the positioner’s feedback loop, it avoids the problem of changes in the positioner loop gain.

7.17 Positioner Guidelines The following have been adapted from a number of guidelines suggested by John Egnew, of Emerson–Fisher. 7.17.1 All Positioner Types • Reducing control valve deadband caused by friction Most control valves without positioners, even those using “low friction” packing material, may demonstrate a 5% deadband. Deadband greater than 1% can produce loop controllability problems. Use of a positioner can reduce deadband caused by friction to less than 1%. • Reducing the effects of frictional stick/slip Stick/slip occurs when a valve first sticks and then jerks (slips) to a new position. Frequently, the stick/slip action causes process variables to overshoot the setpoint. Control action reverses the signal to the control valve and the stick/slip action is repeated in the opposite direction. A positioner’s stem feedback can reduce stick/slip effects and maintain the process variable closer to the setpoint. • Split ranged control elements Processes requiring extended flow rangeability may use two control valves. The first valve operates over the first half of the range. When the first valve is nearly 100% open, the control action begins to open the second valve. This is called control valve split ranging. When split ranging is performed on the pneumatic signal to the valve actuators, positioners on each valve are used to achieve full valve travel over the reduced input range. Most applications need a predictable overlap region in the middle of the signal to avoid a zone of no control. Positioners provide the accuracy to ensure the correct overlap exists. • Increasing the seating force and improving shutoff A positioner will drive the output pressure to either zero or full supply pressure whenever the valve reaches a physical travel stop.

164  •   The Concise Valve Handbook

An air-to-close actuator can use the full supply pressure to provide greater valve seat loading. Completely removing the air signal from air-to-open actuators allows the full force of the spring to load the valve seat. • Double-acting actuators All double-acting piston actuators must have a positioner because the pressure on both sides of the piston must be precisely ­controlled. This can only be accomplished with the stem fed back in a ­positioner. 7.17.2 Digital Positioners Intelligent digital positioners can extend capabilities and benefits of traditional analog positioners. Using digital communications, such as HART, Profibus, and Foundation Fieldbus, digital positioners provide the features listed earlier, but can also provide the following additional benefits. • Automatic calibration Digital positioners can perform remote and automatic zero and span calibration in a few minutes—a task that can take a few hours with non-digital positioners. • Characterization The output signal can be characterized to match the system to achieve a linear process with constant gain. The main advantage over cam linearization is that it is performed on the output signal, not the feedback from the valve stem. • Digital noise filtering While filters should be applied carefully, where appropriate, a digital filter time constant can be applied to minimize the effects of excessive process noise. Users should remember filters add to process response times. Applying a filter will likely require retuning the control loop. • Alarm generation Users can assign positioner-based alarms, such as valve travel deviation from the input signal, travel beyond a certain point, and others. These alarms can be displayed on operator graphics. • Maintenance-related data Digital positioners can track valve reversal and total stem travel data that can be correlated with time and actual maintenance events to improve predictive maintenance forecasting.

Valve Actuators and Positioners   •  165

• Valve stroke speed control In applications where hydraulic hammering might occur a digital positioner can be used to slow the valve stroke. • Limiting the valve travel Digital positioners allow the application of travel limits in cases where a valve should never reach a fully closed position. • Automation of installed valve performance testing Digital positioners communicate well with control valve performance software. Following a pre-defined test, data curves and calculated results can be compared with previous tests to help determine whether a control valve requires maintenance. Being able to compare control valve performance can save time and money during planned outages by focusing maintenance activities on the control valves needing maintenance. • Less susceptibility to vibration influences The solid-state electronics in digital positioners provide a device with few moving parts and help maintain positioner performance in high-vibration installations.

CHAPTER 8

Valve Testing and Diagnostics In Chapter 1, we looked at the process gain: the percentage change in the PV (Process Variable) as a result of a given percentage change in the PD (Process Demand). Another consideration is not by how much the PV changes, but how it changes. A step change of the output will not necessarily result in the same step change in the final control element. This may be as a result of a number of factors that are usually described under the terms deadband and hysteresis.

8.1 Deadband and Hysteresis Deadband is defined as the range through which an input can be varied without initiating an observable response. In a mechanical system, deadband may be, for example, as a result of mechanical play within a gear-train (Figure 8.1). As the input is increased from point A, there will be no change in output until the ‘slack’ is taken up and point B is reached—the extent of the deadband. A well-engineered valve should respond to signals of 1% or less to provide effective reduction in process variability. However, it is not uncommon for some valves to exhibit deadband as great as 5% or more. In a recent plant audit, 30% of the valves had deadbands in excess of 4%. Over 65% of the loops audited had deadbands greater than 2%. In a linear system, the output will now increase proportionally while the input increases. However, when the input (points C to D) is reversed, the output will again not change until the slack or deadband is taken up. Imagine now, a gear-train system in which there is no play, but in which there is an element of elasticity within the gears. As shown in Figure 8.2, the system is no longer linear, and the output does not increase proportionally with the input. This non-linearity is because of energy absorption that

168  •   The Concise Valve Handbook C

Deadband

A

Output

Output

D

Hysteretic error

B Input

Output

Figure 8.1.  Deadband as a result of mechanical play within a geartrain.

Input

Figure 8.2.  An illustration of hysteretic error.

Hysteresis

Input

Figure 8.3.  Hysteresis: A combination of deadband and hysteretic error.

appears as heat. This is called the hysteretic error. When the hysteretic error is summed with deadband (Figure 8.3), it is called hysteresis. Hysteretic error normally manifests itself in a mechanical system that is subject to a cyclic mechanical force or in a magnetic system that is subject to a cyclic magnetizing force. In most pure electronic systems, the hysteretic error can be effectively ignored, and thus the deadband and hysteresis are one and the same. It should also be noted that, in most mechanical-based systems, especially valves, the hysteretic error is referred to as hysteresis, and reference is, therefore, made to ‘deadband and hysteresis.’ Friction is a major cause of deadband in control valves. Rotary valves are often very susceptible to friction caused by the high seat loads required

Valve Testing and Diagnostics   •  169

to obtain shut-off with some seal designs. Because of the high seal friction and poor drive-train stiffness, the valve shaft winds up and does not translate motion to the control element. As a result, an improperly designed rotary valve can exhibit significant deadband that clearly has a detrimental effect on process variability.

8.2 Testing Procedures* *Some of the following information has been gleaned from notes produced by Michael Brown from Michael Brown Control Engineering CC using the Protuner Loop Tuning software. 8.2.1 Deadband and Hysteresis In order to determine the amount of hysteresis plus deadband in a control valve, it is necessary to perform a series of step changes to the input and to observe the changes in the stem position. The method illustrated here is carried out using a Protuner. The results are shown in Figure 8.4 and are the result of two steps up, three down and one up. The first step up overcomes the effects of any hysteresis or deadband to obtain a starting point for the testing.

Equal input step changes

Valve stem position indication

% offset in valve position

Figure 8.4.  Determining the effects of hysteresis/deadband as a result of an input change of two steps up, three down and one up (courtesy Michael Brown Control Engineering).

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The generally acceptable limits are: • • • •

Spring and diaphragm Spring and diaphragm with positioner Piston with positioner Variable speed drive

3% 1% 1% 1%

8.2.2 Stick-Slip ‘Stick-slip’ response in a control valve is the result of a difference in the static and sliding friction in the valve assembly. First, in order to move the valve, the applied force is increased to a level where the static friction is overcome. This is the ‘stick’ phase. Once the static frictional forces are overcome, the sliding frictional forces are much smaller than the static frictional forces and the valve ‘pop’ to a new position. This is the ‘slip’ phase. Stick-slip in the final control element can result in a continuous limit cycle that can destabilize the process. The effects of stick-slip, without hysteresis and deadband, are shown in Figure 8.5. It should be noted that stick-slip is a completely different phenomenon to hysteresis and deadband. 8.2.3 Non-Linearity In an ideal control loop, there should be a linear relationship between the PD and the PV—in other words, the process gain should be constant.

Slip Stick Figure 8.5.  The effects of stick-slip, without hysteresis and deadband (courtesy Michael Brown Control Engineering).

Valve Testing and Diagnostics   •  171 50.0

37.5

PV 25.0

12.5

PD

0.0

Figure 8.6.  As the PD increases in regular step, the PV increases in a series of steps that gradually become smaller, showing a marked non-linearity that is typical of an ­oversized  valve.

Testing for non-linearity is fairly easily accomplished by applying a series of equal-interval steps to the PD as shown in Figure 8.6. Here, PV is shown as a series of steps that gradually become smaller as the PD increases and shows a marked non-linearity that is typical of an oversized valve. Or, is it? The assumption here, of course, is that the PV is a linear measurement. In the vast majority of cases, this would undoubtedly be true, but it need not necessarily be so. 8.2.4 Testing a Complete Assembly It should be appreciated that, in testing a complete final control element, it is not only seal and packing friction within the valve itself that contributes to deadband and hysteresis. Other factors include: • • • • • •

inadequate air supply; loose or worn linkages in actuator connector; defective or improper calibration of I/P converter; loose or worn linkages in positioner; defective or improper calibration of positioner; and undersized actuator.

In order to test a complete pneumatically operated final control assembly, an accurate valve stem position transmission should be connected to

172  •   The Concise Valve Handbook

1

140 kPa

2 3

PD I/P 4 – 20 mA converter 20 – 100 kPa

4 20 – 100 kPa

Positioner

140 kPa

5

Mechanical linkage

Valve stem 6 position 4 – 20 mA transmitter

Figure 8.7.  Testing connections for a complete pneumatically operated final control assembly.

the valve. In addition, pressure transducers should be connected to the airlines, via quick-disconnect pneumatic fittings, at points 3, 4, and 5 as shown in Figure 8.7. All six measuring points should then be connected to a six channel analyzer/recorder. Two series of tests should be carried out. The first test should involve several steps in both directions, with a minimum of two steps up, three down, and one up again. The valve should then be slowly ramped up and down to test for stick-slip. In evaluating the valve-installed dynamic operation, you should look for the following points: • Examine the change in the supply pressure during changes in the valve position. If it drops during position changes, the supply line size should be increased. • Determine both the repeatability and linearity of the I/P converter by recording both its input and output. • Compare the PD signal (input to the I/P converter) with the valve stem position to determine linearity, repeatability, and the response time constant. • Determine whether the valve is stroking correctly. In normal modulating control applications, the valve should never be operated either fully open or at flows of less than 20% of maximum. The evaluations should, therefore, be carried out within these two limits. One of the effects of an undersized actuator is shown in Figure 8.8. If the actuator is unable to cope adequately with the static friction (stiction),

Valve Testing and Diagnostics   •  173

Equal input step changes

Valve stem position indication

Overshoot indicates negative hysterisis

Figure 8.8.  Negative hysteresis: one of the effects of an undersized actuator (courtesy Michael Brown Control Engineering).

then, on reversal, the valve is effectively stuck. As a result, the valve positioner starts to force excess air pressure into the actuator. Once it moves and the static friction is overcome, overshoot occurs. This characteristic is likely to result in continuous cycling.

8.3 Online Diagnostics The tests outlined above have been used for many years to perform analyses on final control elements as a part of routine maintenance. Now, however, using modern Fieldbus communication systems, together with intelligent Smart positioners, a wide variety of information can be monitored on a continuous basis to provide a real-time continuous process and instrument diagnostic system. At the simpler level, both the total valve stem travel (travel accumulation) and the number of stem travel reversals (cycles) can be monitored in order to determine ‘usage.’ Further, should a problem show up, users can define how the instrument reacts to the problem. For example, if the pressure sensor fails, should the instrument be shut down? Users can also select which problems will cause the instrument to shut down. Such indicators maybe reported as alerts to give an instant indication of any problem with the instrument valve or process. Advanced diagnostics also include signature analysis that allows users to determine the valve/actuator friction, Bench Set, spring rate, and seat load.

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8.3.1 Signature Analysis In the ‘valve signature’ plot (Figure 8.9), the actuator pressure (input) is plotted on the y-axis, while the travel (output) is plotted along the x-axis. Examination of this plot shows spikes (change of slope) at each end of the curve, which verifies that a solid stop had been reached at both ends of travel. Next, the opening and closing lines should be parallel and linear throughout the full stroke. The separation of these lines is the result of the friction band—the higher the friction, the wider the separation. Because friction opposes motion, in both directions, the separation between these lines is actually double the friction (friction opposing the up stroke plus friction opposing the down stroke). The primary source of friction on a good valve is the valve packing. Packing materials that have a high coefficient of friction, such as graphite, will produce a greater amount of friction, and thus a wider bandwidth than the low-coefficient materials, such as PTFE. Regardless of the packing material, the separation (friction band) should remain constant throughout the full travel. The slope indicates that the actuator contains an opposing spring. If there were no spring, the opening and closing lines would be nearly flat

Top stop

1.4 1.2 1.0 0.8

2 × Friction Closed

Actuator pressure (bar)

1.6

ning

ing

Clos

Ope

0.6 0.4 Bottom stop

0.2 0 10

0

10 20 Travel (mm)

30

Figure 8.9.  Plotting the ‘valve signature’ plot with the actuator pressure plotted on the y-axis and the travel plotted along the x-axis. The separation between the opening (red) and closing (blue) lines is the result of the friction band (courtesy Fisher-Emerson).

Valve Testing and Diagnostics   •  175

(horizontal), and thus the actuator spring and the effective area of the actuator diaphragm govern the slope’s angle. Examining this data allows a full analysis to be performed. For example, by looking at individual pairings of adjacent upstroke/downstroke data points between 10% and 90% of the travel range, it is possible to derive the minimum, maximum, and average friction values. The minimum friction value should never be less than 25% of the expected friction value (20% if PTFE packing), while the maximum value should never exceed 100% of the expected value. Fairly obviously, there should not be a large difference between the minimum and maximum values, which is the result of calculating the difference in actuator pressure between the upstroke and downstroke, times the effective area of the actuator, divided by two. Figure 8.10 shows that the packing friction is approximately twice that of the previous example. Typically, this might be due to errantly over-tightening the packing, resulting in the excessive friction. Because the total amount of available actuator force is limited via the installed spring, diaphragm area, and air supply, any additional force required to travel the valve through any excess friction must come from some  area within this limitation. The only force available is the one reserved for seat loading, and thus, any increase in friction will diminish seat load.

Top stop

1.4

2 × Friction

1.2

Closed

Actuator pressure (bar)

1.6

1.0 0.8

ning

Ope

ing

Clos

0.6 0.4 0.2

Bottom stop

0 10

0

10 20 Travel (mm)

30

Figure 8.10.  The packing friction is approximately twice that of the previous example. Typically, this might be due to errantly over-tightening the packing, resulting in the excessive friction (courtesy Fisher–Emerson).

176  •   The Concise Valve Handbook Top stop

1.4 1.2

Closed

Actuator pressure (bar)

1.6

1.0 0.8

ning

g

sin

Ope

Clo

0.6 0.4 0.2

Bottom stop

0 10

0

10 20 Travel (mm)

30

Figure 8.11.  An example of a valve signature showing several revealed faults (courtesy Fisher–Emerson).

Figure 8.11 shows an example of a valve signature in which the large increase in friction, as the valve travels towards the open position, is probably due to some form of galling.

8.4 Electronic Torque Monitoring Generally, we have confined ourselves to diagnostics within a pneumatic system. However, as indicated in Chapter 5, modern electrical actuators can monitor both torque and position, allowing data relating to these parameters to be used to compare the footprint torque characteristic of a valve during installation to subsequent torque profiles. As shown in Figure 8.12, the valve position (travel) is plotted on the x-axis, while the torque demand is plotted on the y-axis. The plot illustrates the opening torque characteristics of a typical wedge gate valve. The significant torque requirements for opening or closing a wedge gate valve occur during the final travel going closed and the initial travel going open. During the remaining portion, the torque demand is essentially due to packing and thread friction. As the valve seats, the hydrostatic force on the closure element (the disk) increases the seating friction, and finally, the wedging effect of the disc into the seat causes a rapid increase in torque until seating is affected. Similarly, when under seating the valve, the disk has to be unwedged, and the hydrostatic force of the differential pressure across the valve has to be overcome as the valve is opened. Once the valve is cracked open and

Valve Testing and Diagnostics   •  177

Torque demand

Galled seat

Stem requires lubrication

Packing too tight

Footprint torque profile

Closed

Valve Position

Open

Figure 8.12.  Opening torque characteristics of a typical wedge gate valve in which the valve position (travel) is plotted on the x-axis and the torque demand is plotted on the y-axis (courtesy Rotork).

the differential pressure has dissipated, then the torque demand drops off significantly. Should, for example, the valve stem packing be over-tightened, then an immediate increase in torque profile would be recorded. Should lubrication on the thread to deteriorate over time, then there would be an incremental increase in overall torque. Alternatively, should the valve seat become galled or deteriorate, then there would be an increase in the unseating torque required.

CHAPTER 9

Valve Maintenance and Repair Valves are dynamic devices that exist in a dynamic environment. • • • •

They have moving parts that wear. They have packing and sealing that age and lose their effectiveness. They are subject to the abrasive and corrosive effects of the fluids. They are subject to fluctuations in pressure and temperature of the process of media. • They are subject to fluctuations in the environmental temperature and vibration. Valves, therefore, need to be maintained, prepared, and sometimes replaced. While maintenance on most valves is limited, repair possibilities are extensive—limited only by economic considerations. Apart from lubrication, maintenance consists mainly of correcting external fluid leakage at the stem or shaft of gate, globe, ball, and butterfly valves. In most cases, stem or shaft leakage can be stopped by tightening the packing flange and nuts, which compresses the packing and forces it tighter against the stem. Any partial disassembly of the valve is termed a repair. Repairs are required when: • there is external leakage at the stem, shaft, or body joints, which cannot be stopped by tightening packing nuts, flanges, or body bolting; • closing the valve does not stop fluid flow; • opening the valve does not allow flow to start; • fluid leakage occurs through the valve shell due to erosion or corrosion;

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• there is excessive deadband/hysteresis; and • there is excessive slip-stick.

9.1 In-Line Repairs The decision to make an in-line repair or send it out to a shop depends on the nature and urgency of the repair and the ease of removal. In general, repair in a shop is preferred over in-line repair. In-line repairs are carried out when: • • • •

repairs must be done promptly; if the valve is large or difficult to handle; the valve is in an awkward position; and the valve is welded into the line.

The extent of in-line repair is limited by the type and design of the valve and whether the line has been pressurized drained.

9.2 Repairs Under Pressure The only in-line repair that can be done while the line is still under pressure is the replacement of stem packing on gate and globe valves having back seats (Figure 9.1).

Back-seat

Back-seat

Figure 9.1.  Although not generally recommended, in-line repair can be carried out while the line is still under pressure, on gate and globe valves having back seats.

Valve Maintenance and Repair   •  181

After tightly back-seating the stem, the valve packing nut is unscrewed, or the gland flange nuts removed and the gland flange and the gland lifted. The old packing can then be removed using mechanical packing pullers or high-pressure water sprays to blow out the packing. When using mechanical pullers, take care not to scratch the stem or wall of the stuffing box because scratches in these areas produce potential leak paths. Although the old packing is usually either a continuous length of material or solid rings, the new packing must be in the form of split rings, so that it can be fitted around the stem, with the splits staggered around the stem to prevent a direct leak path. After installing the new packing, the packing or gland flange nuts are reattached and tightened. The stem can then be carefully unseated from the back seat and the packing or gland flange nuts tightened further to stop any leakage. Repacking a valve while it is under pressure is not a universally accepted procedure. Some manufacturers recommend against it, stating that the back seat should be used only as a temporary measure to minimize leakage until repacking can be done in a depressurized condition. Furthermore, many technicians refuse to carry out on-line repacking, particularly on valves carrying steam or other high-pressure or high-temperature fluids.

9.3 Repairs on Drained Systems If the pipeline and valve have been drained, many valves allow the valve bonnet cover to be removed, allowing removal of the flow control element and exposing the body seating services.

9.4 Packing Replacement Because the stem or shaft can be removed from the bonnet or cover, replacing the stem packing is greatly facilitated, with the new packing either continuous material or solid rings. Whenever a valve is disassembled for any repair, it is considered good practice to replace packing, seals, and gaskets.

9.5 Replacing or Refinishing Seat Rings For top-entry valves, the removal and replacement of seat rings is often easily accomplished, especially if they are screwed in. However, special spanner wrenches may be required.

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If the seat rings are tack welded to the body, the tack weld must be cut away. In the case of fully welded-in seat rings or integral seating surfaces, these can be refinished, still with the valve in place, using special machinery that clamps onto the valve. It may be necessary to perform further grinding or lapping of the seat rings to get a good leak-tight fit.

9.6 Other In-Line Repairs The balls in a ball valve are not easily refinished and are usually replaced. However, this is seldom necessary because the soft seat rings experience almost all the wear. Diaphragm valve diaphragms are easily replaced and their fit with their seating services is not a concern.

9.7 In-Line Post-Repair Procedures After an in-line repair, the valve should be pressure-tested to check the integrity of the body–bonnet or body–cover joint and seat tightness. Because this would involve all or at least part of the pipeline, it may not be possible to test to the pressures that would be used to test the valve alone (normally 1.5 times the Maximum Working Pressure (MWP) of the valve). Consequently, the recommended practice is to test at 1.5 times the design pressure of the pipeline. Checking for seat leakage is also difficult after an in-line repair because the seating services are not visible. As a result, seat leakage must be checked at a point downstream of the valve.

9.8 Shop Repairs Because there is much better control in a shop, the quality is better and the valve can be more effectively tested for shell and seat tightness. Furthermore, if the valve can be replaced with a spare one, the downtime of the piping system may be less than with an in-line repair and the repaired valve can then be become a spare. During periodic shutdowns, it is a common practice to remove all the large bore valves and send them to the shops for disassembly, inspection, and if necessary, repair. For economic reasons, small bore piping

Valve Maintenance and Repair   •  183

systems, including their valves, are often scrapped and replaced, rather than repaired. Work performed in a valve shop should be controlled by a valve repair specification that defines the repairs permitted and the quality of work required. These specifications address the following topics: 1. Disassembly and cleaning. • The match marking of parts to ensure proper reassembly and the methods permitted for cleaning without damaging them. 2. Inspection. • Define the visual and dimensional standards to be used to check cleaned parts. 3. Evaluation of inspection results. • Defining the criteria for the rejection of inspected parts and the valve as a whole. • A limit is usually set on the allowable cost of repair of a valve. • A percentage of the cost of a new valve is usually established. • If the cost of repair exceeds the percentage, the valve is scrapped. 4. Permitted repairs. • Define the repairs allowed for specific valve parts. • Should include requirements for the welding of parts, refinishing of weld overlay of seating services, machining of stems and shafts, and machining of flange faces. 5. Reassembly. • Define the need to replace packing gaskets, seals, and bolting and their requirements. • Specify dimensional requirements for reassembled valves. 6. Testing. • The need to pressure test each repaired valve is stated and the specific testing requirements are defined. Such standards (e.g., ANSI) define the different tests required, test media, pressure duration, and the allowable leakage. 7. Preparation for shipment. • In painting and tagging requirements, the records necessary to document the work performed or established.

9.9 Actuator Bench Set Although definitions vary, at its simplest, the term ‘Bench Set’ entails selecting the correct actuator spring. More formally, it is ‘the calibration

184  •   The Concise Valve Handbook

of the actuator spring range of a control valve to account for the in-service process forces.’ For a given actuator/valve assembly, this involves selecting the optimum actuator spring characteristics to achieve a mechanical force equal to, or greater than, all the forces acting against the valve throughout its rated travel range while in service. The forces acting against the valve include: process forces, frictional forces, seating force, and forces due to special assemblies (such as those with multiple piston rings). Bench Set is, thus, a predetermined value that is established during the actuator sizing procedure. The name, Bench Set, stems from the fact that this test is usually performed on a workbench in the instrument shop, prior to placing the valve into service. Factors that influence the Bench Set span are: actuator size, spring characteristics, and rated valve travel. For a given actuator/valve assembly, the valve travel and diaphragm size are known, and thus, the Bench Set establishes the required spring characteristics: the spring compression ratio—sometimes referred to as the spring windup. Often, setting the spring windup is carried out with the actuator disconnected from the valve so as not to introduce any forces, particularly frictional forces. Typically, valves use two stops: one being the valve seat and the other the top actuator stop. Since, with the actuator removed from the valve, there is no way of knowing where the bottom stop (seat) will be, the only common reference point for establishing the Bench Set would be the upper stop of the actuator. With this in mind, the Bench Set spring windup and span are established from the upper actuator stop down to the rated travel of the valve on which the actuator will be installed.

9.10 Spring Calculations But first, a recap on a few basics. Pressure is defined as the ‘force per unit area’ with units of psi, mm Hg, bar, and kPa. One of the most well-known is lbs/inch2. So, the question if this is correct, what about kg/m2? The answer is no! Why? Because the term ‘pound’ is a unit of force, while the term kilogram is unit of mass. In reality, the SI unit of force is the Newton, which is defined as the force that gives a 1 kg mass an acceleration of 1 m/s2. Consequently, SI pressure is defined as N/m2, or more commonly, the Pascal (Pa). Practically, because the Pascal is so small, use is made of the kilopascal (kPa) in which 100 kPa = 1 bar.

Valve Maintenance and Repair   •  185

OP EN

Now, some idea of the parameter’s values involved can be gained by examining Figure 9.2: a nominal 50-mm-diameter valve. Assume a process inlet pressure (P1) of 3 bar, and in its fully open position, an outlet pressure (P2) of 2 bar, resulting in a ∆P of 1 bar (10 N/ cm2). For a 50 mm valve, the port area is 19.6 cm2, and thus, there is a force of 19.6 ×10 = 196 N acting along the valve stem, which is tending to open the valve, and which is assumed to be constant over the full length of travel. To start closing the valve, therefore, the actuator would need to exert a downward force of 196 N at the minimum operating air input of 20 kPa (2 N/cm2), necessitating a diaphragm area of 196/2 = 98.2 cm2. Assuming that the actuator is ranged correctly, then, at the full instrument air input of 100 kPa (10 N/cm2), the valve should be fully closed. At that point, the closing force is 98.2 ×10 = 982 N; an excess of 982–196 = 786 N above that required to close the valve (Figure 9.3). The absorption of this surplus force is the function of the ‘ranging’ spring. With, for example, a valve having a full travel of 1.75 cm, the spring would need to have a compression rate of 786/1.75 = 450 N/cm. In this manner, it is possible to choose the diaphragm size and the spring compression ratio from a number of standard values in order to satisfy the requirements of the valve, to within around 3 to 4 kPa. From the foregoing, it is important to note that, before reusing a control valve in a different application, mixing and matching valves and

P2 = 200 kPa

P1 = 300 kPa ∆P = 100 kPa

Figure 9.2.  Forces developed on a nominal 50-mm valve plug.

196 N

OP EN

186  •   The Concise Valve Handbook

982 N

Figure 9.3.  Unbalance of forces.

actuators, adding or removing a positioner, changing components that influence hysteresis, or adjusting spring compression, it is best to discuss the planned changes with the original valve manufacturer.

CHAPTER 10

Safety Relief Valves The primary function of a pressure relief device is to prevent overpressure by automatically opening and releasing a volume of fluid (gas or liquid) from within the vessel when a predetermined maximum pressure is reached. Acting as a ‘last resort,’ these fully mechanical devices are designed to open when an overpressure situation occurs within a process pressure system, protecting not only life, but safeguarding the investment and plant itself. Such devices have been around since the 1600s and may be classified as either reclosing or non-reclosing (Figure 10.1). Pressure safety relief valves should be taken very seriously. Manufactured from castings, they may not look very sophisticated, but in their design, accuracy, and function, they resemble a delicate instrument, while at the same time, performing an essential role. Self-contained and self-operating, they respond to system conditions and prevent catastrophic failure when other instruments and control systems fail to control process limits adequately.

10.1 History It is usually supposed that the Frenchman Papin was the inventor of the safety valve, which he first applied about in 1682 to a digester. The valve was kept closed by means of a lever and movable weight; sliding the weight along the lever enabled Papin to keep the valve in place and regulate the steam pressure (Figure 10.2). However, it now appears that safety valves were already in use some 50 years earlier by the German Glauber and that Papin only improved on Glauber’s device. In Glauber’s treatise on philosophical furnaces,

188  •   The Concise Valve Handbook Pressure Relief Devices

Non-Reclosing Pressure Relief Devices

Pressure Relief Valves

Weight Loaded Pressure/Vacuum Relief Valve

Spring Loaded Pressure Relief Valve

Pilot Operated Pressure Relief Valve

Rupture Disc

Buckling/Shear Pin

Safety Valve

Conventional

Conventional

Scored Tension

Relief Valve

Balanced

Composite

Reverse Action

Safety Relief Valve

Graphite

Figure 10.1.  Family of pressure relief devices, classified as either reclosing or non-reclosing.

Figure 10.2.  Papin’s safety valve was kept closed by means of a lever and movable weight. Sliding the weight along the lever kept the valve in place and regulated the steam pressure.

translated into English in 1651, he describes how he prevented retorts and stills from bursting from an excessive pressure. A conical valve was fitted, being ground air-tight to its seat, and loaded with a ‘cap of lead,’ so that when the vapor became too ‘high,’ it slightly raised the valve and a portion escaped. The valve then closed again on itself, ‘being pressed down by the loaded cap and so kept closed.’

Safety Relief Valves   •  189

The idea was followed up by others, and we find in the art of distillation, by John French, published soon afterward in London, the following concerning the action of such safety valves: Upon the top of a stubble (valve) there may be fastened some lead, that if the spirit be too strong, it will only heave up the stubble and let it fall down. It should be realized that the word ‘steam’ was unknown at the time and was only coined sometime later. In its place, we find the words ‘vapor,’ ‘spirit,’ ‘smoke,’ and even, ‘ghost.’ In the early 1800s, there were literally thousands of boiler explosions in the United States and Europe. However, there were no legal codes for boilers in the United States. During the five years between 1905 and 1911, there were 1,700 boiler explosions resulting in 1,300 deaths in the New England region of the United States alone. Boiler failure in Brockton, Massachusetts, on March 10, 1905, at the Brockton Shoe Factory resulted in 58 deaths and 117 injuries, and completely leveled the factory. In 1906, Massachusetts established a five-man Board of Boiler Rules, whose charge was to write a boiler law for the state. This was published in 1908. And, in 1911, the State of Ohio enacted a boiler law. By 1911, the year in which the ASME Council appointed a committee to formulate a boiler code, there were laws and regulations in effect in 10 states and 19 metropolitan areas. The individual state requirements differed greatly from one another, so a boiler built in one state could not be installed in another state. Consequently, the ASME Council established the Boiler Code Committee to prepare a standard that could be accepted by all states. The committee’s mission was to formulate a standard specification for the construction of steam boilers and other ­pressure vessels. In 1916, the ASME Council approved the formation of the Conference Committee to provide technical input, as it sees fit, to the additions and revisions to the Boiler and Pressure Vessel Code. All the provinces in Canada, 48 of the 50 states of the United States, and various regulatory agencies around the world have adopted, by law or regulation, various sections of the Boiler and Pressure Vessel Code. The American Society of Mechanical Engineers was asked to formulate a design code. The boiler and pressure vessel committee was formed, and hence, the ASME Section 1 for fired vessels was formulated, becoming a mandatory requirement for all states that recognized the need for

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regulation. The sole purpose of a pressure-relieving device (safety relief valve) is to protect life and property.

10.2 Definitions The terms ‘safety valve’ and ‘safety relief valve’ are generic terms that describe a variety of pressure relief devices designed to prevent excessive internal fluid pressure build-up. One of the first problems encountered in the field of pressure safety relief valves lies in the differences in terminology used between the United States and Europe. One of the most important is that a valve referred to as a ‘safety valve’ in Europe is referred to as a ‘safety relief valve’ or ‘pressure relief valve’ in the United States. Furthermore, the term ‘safety valve’ in the United States generally refers specifically to the full-lift type of safety valve used in Europe. The European standards (BS 6759 and DIN 3320) provide the following definition: Safety valve: A valve that automatically discharges a certified amount of fluid to prevent a predetermined safe pressure being exceeded. When normal pressure conditions have been restored, the valve will close and prevent further fluid flow. In the United States, the ASME/ANSI PTC25.3 standards define the following generic terms: Pressure relief valve is a generic term that includes safety valves, relief valves, and safety relief valves. In essence, a pressure relief valve describes any spring-loaded device designed to open and relieve excess pressure—re-closing after normal conditions have been restored to prevent further fluid flow. All three varieties are similar in design and operation, but have different applications. Safety valves are primarily used with compressible gases, particularly for steam and air services, and are characterized by a rapid-opening ‘pop’ action. Relief valves are commonly used in liquid systems, especially for lower capacities and thermal expansion duty, and are characterized by a gradual lift, generally proportional to the increase in pressure over opening pressure. Safety relief valves may be used for either liquid or compressible fluid applications and will perform as a safety valve when used in a compressible gas system (characterized by a ‘pop’ action) and as a relief valve when used in liquid systems (characterized by a proportional opening action).

Safety Relief Valves   •  191

10.3 Weight-Loaded Pressure/Vacuum Relief Valves Direct-acting, weight-loaded valves are the simplest and least complex type of pressure relief devices. Commonly referred to as weighted pallet valves, they are frequently used on storage vessels requiring breather vents. As illustrated in Figure 10.3, the weight of the seat assembly (or pallet) keeps the valve closed until the pressure acting on the underside equals this weight. Normally designed for set pressures of less than 0.14 bar (2 psig), weight-loaded valves are not ASME-coded. A 30 cm (12 inch) valve (commonly used on large storage tanks) typically has a nozzle area of approximately 580 cm2 (90 inch2). In order to obtain a set pressure of only 0.07 bar (1.0 psig), a mass of 40 kg (90 lb) weight would be required. Consequently, large valves are usually limited to pressures of 0.03 bar (0.5 psig) or less because of the excessive weight required to obtain the set pressure. When associated with tank breathing (1 to 4 in WC (0.04 to 0.15 psi)), weight-loaded valves are often referred to as a ‘conservation valves’ and comprise a combination vacuum plus a pressure valve to provide IN- and OUT-breathing (Figure 10.4). At these very low pressures, the relief valve/vent will almost always be weight-loaded since it is not economically feasible to manufacture a spring for such low pressures. The main advantages of weight-loaded valves are their low cost, their simple operation, and the ability to set very low pressures (down to 0.5 ounce/in2).

Force due to seat weight

Stem guide assembly Seat stem Seat assembly/pallet Nozzle Medium pressure

Figure 10.3.  The weight of the seat assembly (or ­pallet) keeps the valve closed until the pressure acting on the underside equals this weight.

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Tank Connection

Air Inlet

Figure 10.4.  When associated with tank breathing (1 to 4 in WC), weight-loaded valves are often referred to as a ‘conservation valves’ and provide IN- and OUT-breathing.

However, there are a number of limitations: • The set pressure is not readily adjustable. • Extremely long simmer and poor tightness. • High overpressure necessary for full lift (100% or more in some cases). • The seat is easily frozen closed at low (cryogenic) temperatures. • The weights required for higher set pressures become prohibitively large. • At higher pressures, many weighted pallet pressure relief valves exhibit oscillation, resulting in seat plate damage. • Accumulation of liquid can occur at the top of the plate.

10.4 Spring-Loaded Relief Valves A relief valve is a direct spring-loaded pressure relief valve actuated by the static pressure upstream of the valve and characterized by a gradual lift proportional to the increase in pressure. Figure 10.5 illustrates a typical relief valve where a spring force opposes the system pressure acting on the valve disc. When the system pressure rises above the level of the spring force, the valve opens. The basic elements comprise a right-angle-pattern valve body with the valve inlet connection, or nozzle, mounted on the pressure-containing

Safety Relief Valves   •  193 Cap Stem Adjusting screw

Spring Bonnet

Seating surface

Disk

Body Nozzle

Figure 10.5.  Spring-loaded pressure relief valve.

system. The outlet connection may be screwed or flanged for connection to a piped discharge system. However, in some applications, such as compressed air systems, the safety valve will not have an outlet connection, and the fluid is vented directly to the atmosphere. The valves have closed bonnets to prevent the release of corrosive, toxic, flammable, or expensive fluids and can be supplied with lifting levers (Figure 10.6), balancing bellows, and soft seats as needed. The ASME Code requires that liquid service relief valves installed after January 01, 1986, have their capacity certified and stamped on the nameplate. The closing force on the disc is provided by a spring, typically made from carbon steel. The amount of compression on the spring is usually adjustable, using the spring adjuster, to alter the pressure at which the disc is lifted off its seat. A relief valve begins to open when the static inlet pressure reaches set pressure. When the static inlet pressure overcomes the spring force, the disc begins to lift off the seat, allowing flow of the liquid. The value of the closing pressure is less than that of the set pressure and will be reached after the blowdown phase is completed. Relief valves usually reach full lift at either 10% or 25% overpressure, depending on the type of valve and trim.

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Spring adjusting screw Test lever

Spring

Stem

Nozzle

Figure 10.6.  The valves have closed bonnets to prevent the release of ­corrosive, toxic, flammable, or ­expensive fluids and can be supplied with lifting levers.

10.5 Applications Relief valves are normally used for incompressible fluids (see Part I of API RP 520).

10.6 Limitations Relief valves should not be used as follows: • In steam, air, gas, or other vapor services. • In services piped to a closed header unless the effects of any constant or variable back-pressure have been accounted for. • As pressure control or bypass valves.

10.7 Safety Valves A safety valve is a direct spring-loaded pressure relief valve that is actuated by the static pressure upstream of the valve and characterized by rapid

Safety Relief Valves   •  195

opening or pop action. When the static inlet pressure reaches the set pressure, it will increase the pressure in the huddling chamber and overcome the spring force on the disc. This will cause the disc to lift and provide full opening at minimal overpressure. The closing pressure will be less than the set pressure and will be reached after the blowdown phase is completed. The spring of a safety valve is usually fully exposed, outside of the valve bonnet to protect it from degradation due to the temperature of the relieving medium. A typical safety valve (Figure 10.7) has a lifting lever for manual opening to ensure the freedom of the working parts. Open bonnet safety valves are not pressure-tight on the downstream side. Standards that govern the design and use of safety valves generally only define the three dimensions that relate to the discharge capacity of the safety valve, namely, the flow (or bore) area, the curtain area, and the discharge (or orifice) area (Figure 10.8). • Flow area: The minimum cross-sectional area between the inlet and the seat, at its narrowest point. The diameter of the flow area is represented by dimension ‘d’ in Figure 10.8.

Spring adjusting screw Test lever

Spring

Stem Lower adjusting ring

Upper adjusting ring

Figure 10.7.  The spring of a safety valve is usually fully exposed and has a lifting lever for manual opening.

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D Curtain area Flow area

L d

Flow Flow

Figure 10.8.  Illustration of the standard defined areas.

• Curtain area: The area of the cylindrical or conical discharge opening between the seating surfaces created by the lift of the disc above the seat. The diameter of the curtain area is represented by dimension ‘d1’ in Figure 10.8.

Flow area =

p ⋅ d2 4

Curtain area = p ⋅ d1 ⋅ L

(10.1) (10.2)

• Discharge area: This is the lesser of the curtain and flow areas, which determines the flow through the valve.

10.8 Basic Operation: Lifting When the inlet static pressure rises above the set pressure of the safety valve, the disc will begin to lift off its seat. However, as soon as the spring starts to compress, the spring force will increase; this means that the pressure would have to continue to rise before any further lift can occur and for there to be any significant flow through the valve. The additional pressure rise required before the safety valve will discharge at its rated capacity is called the overpressure. The allowable overpressure depends on the standards being followed and the particular application. For compressible fluids, this is normally between 3% and 10%, and for liquids, between 10% and 25%. In order to achieve full opening from this small overpressure, the disc arrangement has to be specially designed to provide rapid opening. This is

Safety Relief Valves   •  197

usually done by placing a shroud, skirt, or hood around the disc (Figures 10.9 and 10.10). The volume contained within this shroud is known as the control or ‘huddling’ chamber. As lift begins (Figure 10.10 (b)) and fluid enters the chamber, a larger area of the shroud is exposed to the fluid pressure. Since the magnitude of the lifting force (F) is proportional to the product of the pressure (P) and the area exposed to the fluid (A), (F = P ¥ A), the opening force is increased. This incremental increase in opening force overcompensates for the increase in spring force, causing rapid opening. At the same time, the shroud reverses the direction of the flow, which provides a reaction force, further enhancing the lift. These combined effects allow the valve to achieve its designed lift within a relatively small percentage overpressure. For compressible fluids, an additional contributory factor is the rapid expansion as the fluid volume increases from a higher to a lower pressure area. This plays a major role in ensuring that the valve opens fully within the small overpressure limit.

Disc Shroud

Huddling chamber

Figure 10.9.  Typical disc and shroud arrangement used on rapid opening safety valves.

(a)

(b)

Figure 10.10.  Operation of a conventional safety valve.

(c)

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For liquids, this effect is more proportional, and subsequently, the overpressure is typically greater; 25% is common.

10.9 Basic Operation: Reseating Once normal operating conditions have been restored, the valve is required to close again, but since the larger area of the disc is still exposed to the fluid, the valve will not close until the pressure has dropped below the original set pressure. The difference between the set pressure and this reseating pressure is known as the ‘blowdown,’ and it is usually specified as a percentage of the set pressure (Figure 10.11). For compressible fluids, the blowdown is usually less than 10%, and for liquids, it can be up to 20%. The design of the shroud must be such that it offers both rapid opening and relatively small blowdown, so that, as soon as a potentially hazardous situation is reached, any overpressure is relieved, but excessive quantities of the fluid are prevented from being discharged. At the same time, it is necessary to ensure that the system pressure is reduced sufficiently to prevent immediate reopening. In order to achieve a significant lifting force without an extremely long blowdown, a ring is threaded around the valve nozzle and positioned to form a huddling chamber with the disc skirt (Figure 10.12). Although this ring, shown as the ‘nozzle ring,’ is commonly called a blowdown ring, its function is also very important for controlling the valve opening. Pressure is generated in the huddling chamber when gas or vapor flows past the seat. This pressure, acting over a larger area than the seat sealing area, increases, and thus creates an instantaneous amplification of the upward force, and the seat disc rapidly lifts off the nozzle. Maximum discharge

100% % lift

Closing

Opening Set pressure

Reseat

10% Blowdown

Pop action

Overpressure 10%

Figure 10.11.  Relationship between pressure and lift for a typical safety valve.

Safety Relief Valves   •  199 Spring

Huddling chamber

Disc Nozzle ring

Nozzle

Figure 10.12.  Blowdown ring is threaded around the valve nozzle and positioned to form a huddling chamber with the disc skirt.

This initial lift of the seat disc is enough to establish 60–75% fullrated flow, driving the seat disc up to the change in momentum and the expansion of the gas can sustain lift. When the blowdown ring is adjusted up (Figure 10.13), the forces required to lift the seat disc off the nozzle occur at a pressure very close to set pressure. The reason for this is that the huddling chamber is restricted and gas flowing into the chamber quickly pressurizes it. However, in this position, the blowdown is long because the pressure between the seat disc skirt and the ring remains high, preventing the seat disc from losing lift until the pressure under the disc reduces to a much lower value.

Increases blowdown Reduces simmer Blowdown ring Decreases blowdown Increases simmmer

Figure 10.13.  When the blowdown ring is adjusted up, the forces required to lift the seat disc occur very close to set pressure and the blowdown is long. When the ring is adjusted down, the seat lift does not occur until the pressure under the seat disc is considerably higher and the blowdown is short.

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When the ring is adjusted down, the forces required to lift the seat disc off the nozzle do not occur until the pressure under the seat disc is considerably higher. This is because the huddling chamber exit area is less restricted and considerably more gas must flow into the chamber to pressurize it. With the ring in this position, the blowdown is short since the pressure between disc holder skirt and ring quickly decreases when the lift of the seat disc is decreased.

10.10 Conventional Safety Relief Valves A conventional safety relief valve is a direct spring-loaded valve whose operational characteristics (opening pressure, closing pressure, and relieving capacity) are directly affected by changes in any back pressure in the discharge system. Conventional safety relief valves can be used in refinery and petrochemical processes that handle flammable, hot, or toxic material. However, the effect of temperature and back-pressure on the set pressure must be considered when using them. 10.10.1 Limitations Conventional safety relief valves should not be used in the following applications: • Where any built-up back pressure exceeds the allowable overpressure. • Where the Cold Differential Test Pressure (CDTP) cannot be reduced to account for the effects of variable back pressure (see API RP 520 Part I). • On ASME Section I steam boiler drums or ASME Section I super-heaters. • As pressure control or bypass valves. 10.10.2 Back-Pressure It is important to note that the total back-pressure is generated from two components: • Superimposed back-pressure. The static pressure that exists on the outlet side of a closed valve.

Safety Relief Valves   •  201

• Built-up back-pressure. The additional pressure generated on the outlet side when the valve is discharging. Subsequently, in a conventional safety valve, only the superimposed back-pressure will affect the opening characteristic and set value, but the combined backpressure will alter the blowdown characteristic and re-seat value. The ASME/ANSI standard makes the further classification that a ­conventional valve has a bonnet that encloses the spring and forms a ­pressure-tight cavity, with the bonnet cavity vented to the discharge side of the valve. If the spring housing is vented to atmosphere, any superimposed back-pressure will still affect the operational characteristics. This is illustrated in Figure 10.14, which shows schematic diagrams of a valve whose spring housing is vented to the discharge side of the valve (an ASME conventional safety relief valve.) Consider the forces acting on the disc (with area AD), the required opening force (equivalent to the product of inlet pressure (PV), and the nozzle area (AN)) is the sum of the spring force (FS) and the force due to the back-pressure (PB) acting on the top and bottom of the disc:

Disk area AD Nozzle area AN

Spring FS PB PB

PB

Vented to discharge

PB PB

Back pressure PB

Nozzle pressure PN

Figure 10.14.  Schematic diagram of a valve with the spring housing vented to the discharge side of the valve.

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PV ⋅ A N = FS + PB ⋅ A D − PB ⋅ ( A D − A N ) 

(10.3)

where: PV = Fluid inlet pressure AN = Nozzle area FS = Spring force PB = Back-pressure which simplifies to:

PV ⋅ A N = FS + PB ⋅ A N 

(10.4)

This shows that any superimposed back-pressure will tend to increase the closing force and the inlet pressure required to lift the disc is greater. Figure 10.15 show a valve whose spring housing is vented to the atmosphere. Here, the required opening force is: PV ⋅ A N = FS − PB ⋅ ( A D − A N ) 



(10.5)

This shows that the superimposed back-pressure acts with the vessel pressure to overcome the spring force and the opening pressure will, as a

Vented to atmosphere

Disk area AD Nozzle area AN

Spring FS

PB

PB

Back pressure PB

Nozzle pressure PN

Figure 10.15.  Schematic diagram of a valve with spring housing vented to the atmosphere.

Safety Relief Valves   •  203

result, be less than expected. In both cases, if a significant superimposed back-pressure exists, its effects on the set pressure need to be considered when designing a safety valve system. Once the valve starts to open, the effects of built-up back-pressure also have to be taken into account. In a conventional valve, with the spring housing vented to the discharge side, once the valve starts to open, the inlet pressure (PV) now becomes the sum of the set pressure, PS, and the overpressure, PO. This modifies the opening force, which now becomes:

( PS + PO ) ⋅ A N = FS + PB ⋅ A N 

(10.6)

PS ⋅ A N = FS + A N ⋅ ( PB − PO ) 

(10.7)

which is simplified to:

where: PS = Set pressure of safety valves AN = Nozzle area FS = Spring force PB = Back-pressure PO = Overpressure From this, it can be seen that, if the backpressure is greater than the overpressure, the valve will tend to close, reducing the flow. This can lead to instability within the system and can result in flutter or chatter of the valve. In general, if a conventional safety valve is used in applications, where there is an excessive built-up back-pressure, it will not perform as expected. According to the API 520 Recommended Practice Guidelines: A conventional pressure relief valve should typically not be used when the built-up backpressure is greater than 10% of the set pressure at 10% overpressure. A higher maximum allowable built-up backpressure may be used for overpressure greater than 10%. The British Standard BS 6759, however, states that the built-up back-pressure should be limited to 12% of the set pressure when the valve is discharging at the certified capacity. For many applications, the back-pressure can be maintained within these limits by carefully sizing of the discharge pipes. If this is not feasible, then it may be necessary to use a ‘balanced’ safety valve.

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10.11 Balanced Safety Relief Valves A balanced safety relief valve is a direct spring-loaded pressure device that incorporates a bellows or piston arrangement to minimize the effect of back-pressure on the operational characteristics. Balanced safety relief valves are normally used in refinery and ­petrochemical process industries that handle flammable, hot, or toxic material, where high back-pressures are present at the valve discharge. This typically occurs where material from the valve is routed to a collection system. They are used as follows: • In gas, vapor, steam, air, or liquid services. • In corrosive service to isolate the spring, bonnet cavity, and discharge side of the valve from process material. • When the discharge from the valves must be piped to remote ­locations. 10.11.1 Limitations Balanced safety relief valves should not be used as follows: • On ASME Section I steam boiler drums or ASME Section I super-heaters. • As pressure control or bypass valves. There are two basic designs of balanced safety relief valves: bellows type and piston type.

10.12 Bellows-Type Balanced Safety Valve In the bellows-type balanced safety valve (Figure 10.16), a bellows with an effective area (AB) equivalent to the nozzle seat area (AN) is attached to the upper surface of the disc and to the spindle guide (Figure 10.17). The bellows prevents back-pressure acting on the upper side of the disc within the area of the bellows. The disc area extending beyond the bellows and the opposing disc area are equal, and so the forces acting on the disc are balanced, and the back-pressure has little effect on the valve opening pressure. The bellows vent allows air to flow freely in and out of the bellows as they expand or contract.

Safety Relief Valves   •  205 Cap Stem Adjusting screw Bonnet Bonnet vent

Spring

Bellows vent

Bellows Disk holder Disk Body Nozzle

Figure 10.16.  Bellows-type balanced safety valve. Spring FS Bonnet vent Bellows vent

Spindle guide

AB Bellows AN

AB Disk

PV

Figure 10.17.  Block schematic of a bellows-­ type balanced safety valve showing force balancing.

Balanced-type valves require vented bonnets. A bellows failure allows process media from the discharge side of the valve to discharge from the bonnet vent. Consider the nature of the process media (e.g., liquid/vapor, toxicity, and flammability) when evaluating the bonnet vent disposition.

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Bonnet vents are typically routed to a drain, a closed system, or atmosphere, depending on the process media involved. Bellows failure is an important concern when using these valves since this may affect the set pressure and capacity of the valve. It is important, therefore, that there is some mechanism for detecting any uncharacteristic fluid flow through the bellows vents. In addition, some bellows balanced safety valves include an auxiliary piston that is used to overcome the effects of back-pressure in the case of bellows failure. This type of safety valve is usually only used in critical applications in the oil and petrochemical industries. In addition to reducing the effects of back-pressure, the bellows also serves to isolate the spindle guide and the spring from the process fluid; this is important when the fluid is corrosive. Since balanced pressure relief valves are typically more expensive than their unbalanced counterparts, they are commonly only used where high-pressure manifolds are unavoidable, or in critical applications where a very precise set pressure or blowdown is required.

10.13 Piston-Type Balanced Safety Valve Although there are several variations of the piston valve, they generally comprise a piston-type disc whose movement is constrained by a vented guide. The area of the top face of the piston, AP, and the nozzle seat area, AN, are designed to be equal. This means that the effective area of both the top and bottom surfaces of the disc exposed to the back-pressure are equal, and therefore, any additional forces are balanced. In addition, the spring bonnet is vented such that the top face of the piston is subjected to atmospheric pressure, as shown in Figure 10.18. By considering the forces acting on the piston, it is evident that this type of valve is no longer affected by any back-pressure:

PV ⋅ A N = FS + PB ⋅ ( A D − A P ) − PB ⋅ ( A D − A N ) 

where: PV = Fluid inlet pressure AN = Nozzle area FS = Spring force PB = Back-pressure AD = Disc area AP = Piston area

(10.8)

Safety Relief Valves   •  207 Spring FS

AP

Bonnet vent

Piston

AD

PB

PB

Vent

PS

Disk PB AN

PB PV

AP = AN

Figure 10.18.  Block schematic of piston-­ type balanced safety valve showing force balancing.

Since AP equals AN, the last two terms of the equation are equal in magnitude and cancel out. This simplifies to:

PV ⋅ A N = FS 

(10.9)

10.13.1 Pilot-Operated Pressure Relief Valve A pilot-operated safety relief valve is a pressure relief valve in which the major relieving device or main valve is combined with, and controlled by, a self-actuated auxiliary pressure relief valve (pilot). As with the springloaded valve, many unique models exist. However, some common design features include: the sensing line, the pilot valve, and the main valve. Depending on the design, the pilot valve (control unit) and the main valve may be mounted on either the same connection or separately. Pilot-operated safety relief valves offer a number of advantages over conventional safety relief valves including: good overpressure and blowdown performance and use in applications where a large relief area and/or high set pressures are required. The pilot is a spring-loaded valve that operates when its inlet static pressure exceeds its set pressure. This causes the main valve to open and close according to the pressure. Process pressure is either vented-off by the pilot valve to open the main valve or applied to the top of the unbalanced piston, diaphragm, or bellows of the main valve to close it.

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The two most commonly used pilot-operated safety valves are the piston and diaphragm types. Figure 10.19 shows a high-pressure pilot-operated valve that uses an unbalanced piston and has an integrally mounted pilot. The piston and seating arrangement in the main valve is designed so that the bottom area of the piston, exposed to the inlet fluid, is less than the top area of the piston. Since both ends of the piston are exposed to the fluid at the same pressure, under normal system operating conditions, the closing force, resulting from the larger top area, is greater than the inlet force. The resultant downward force holds the piston firmly on its seat (Figure 10.20). If the inlet pressure rises, the net closing force on the piston also increases, ensuring that a tight shut-off is continuously maintained. However, when the inlet pressure reaches the set pressure, the pilot valve will pop open to release the fluid pressure above the piston. With much less fluid pressure acting on the upper surface of the piston, the inlet pressure generates a net upward force, and the piston will leave its seat. This causes the main valve to pop open, allowing the process fluid to be discharged. When the inlet pressure has been sufficiently reduced, the pilot valve will reclose, preventing the further release of fluid from the top of the piston, thereby re-establishing the net downward force and causing the piston to re-seat. Set pressure adjustment screw Pilot piston External blowdown adjustment Seat

Optional pilot filter

Piston Seat

Outlet

Pilot supply line

Internal pressure pickup Main valve Inlet

Figure 10.19.  High-pressure pilot-operated valve incorporating an unbalanced piston and an integrally mounted pilot.

Safety Relief Valves   •  209

Cover Spring Soft disc Sliding rings

Guide Disc holder

Disc

Metal disc

Body

Nozzle

Figure 10.20.  Alternative seating arrangements available for a pilot-operated piston-type safety relief valve.

Figure 10.21 shows a diaphragm-type pilot-operated valve that is, typically, only available for use in low-pressure applications and produces a proportional type action—characteristic of relief valves used in liquid systems.

Figure 10.21.  Low-pressure diaphragm-type pilot-operated valve.

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Pilot-operated safety valves offer good overpressure and blowdown performance (a blowdown of 2% is attainable). For this reason, they are used where a narrow margin is required between the set pressure and the system operating pressure. Pilot-operated valves are also available in much larger sizes, making them the preferred type of safety valve for larger capacities. One of the main concerns with pilot-operated safety valves is that the small bore, pilot-connecting pipes are susceptible to blockage by foreign matter, or due to the collection of condensate in these pipes. This can lead to the failure of the valve, either in the open or closed position, depending on where the blockage occurs. The British Standard BS 6759 states that all pilot-operated safety valves should have at least two independent pilot devices, which are connected individually and arranged such that failure of either of the pilot will still enable the safety valve to continue to operate effectively. 10.13.2 Applications Pilot-operated safety relief valves are generally used as follows: • Where a large relief area and/or high set pressures are required, since pilot-operated valves can usually be set to the full rating of the inlet flange. • Where a low differential exists between the normal vessel operating pressure and the set pressure of the valves. • On large low-pressure storage tanks (see API Standard 620). • Where very short blowdown is required. • Where back-pressure is very high and balanced design is required, since pilot-operated valves with the pilots either vented to the atmosphere or internally balanced are inherently balanced by design. • Where process conditions require sensing of pressure at one location and relief of fluid at another location. • Where inlet or outlet piping frictional pressure losses are high. • Where in-situ, in-service, set pressure verification is desired. 10.13.3 Limitations Pilot-operated safety relief valves are not generally used as follows: • In service where fluid is dirty, unless special provisions are taken (such as filters, sense line purging, etc.)

Safety Relief Valves   •  211

• In viscous liquid service, as pilot-operated valve operating times will increase markedly due to flow of viscous liquids through relatively small passages within the pilot. • With vapors that will polymerize in the valves. • In service where the temperature exceeds the safe limits for the diaphragms, seals, or O-rings selected. • Where chemical compatibility of the lading fluid with the diaphragms, seals, or O-rings of the valves is questionable. • Where corrosion build-up can impede the actuation of the pilot.

10.14 Non-Reclosing Pressure Relief Devices While pressure relief valves are designed to automatically reset to their previous condition once the overpressure condition has ceased, non-reclosing pressure relief devices generally require the device to be replaced. In essence, non-reclosing pressure relief devices are available in two forms: buckling/shear pin devices and rupture discs.

10.14.1 Buckling/shear pin devices 10.14.1.1 Shear-Pin Safety Valve The shear-pin safety valve is similar in construction to a spring-loaded valve, but instead of a spring, use is made of a shear-pin (Figure 10.22). As shown, the valve spindle slides in a guide, with both drilled to accept the shear-pin. When the set pressure is reached, the force exerted on the disc equals the failure shear force of the pin. The use of two pins can extend the set pressure range up to 5,000 psi. The shear-pin failure force is dictated by the pin’s material strength and cross-sectional area. Because of the lack of precision in determining the shear failure rate, this form of device is not ASME-approved.

10.14.1.2 Buckling Pin Safety Valve The basic buckling pin valve (Figure 10.23) comprises a pin of a ­precise length that holds a piston on its seat. The pin ends are restrained for ­precise, repeatable operation.

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Spindle guide Shear pin Spindle

Figure 10.22.  Similar in construction to a spring-loaded valve, but using a shear-pin in place of a spring.

Buckling pin

Figure 10.23.  The basic buckling pin valve comprises a pin of a precise length that holds a piston on its seat. As the pressure increases and the axial force on the pin subsequently also increases, the pin will buckle.

As the pressure increases and the axial force on the pin subsequently also increases, the pin will buckle. This ‘buckling’ point is based on Euler’s law of Compressed Columns that states:

F≈

E ⋅ d4 L2



(10.10)

Safety Relief Valves   •  213

where: F = axial force causing the pin to buckle E = pin modulus of elasticity d = pin diameter L = pin length When the set pressure is reached, the pin buckles and the piston rapidly moves off its seat to relieve the pressure. The pin has only two stable conditions: straight or buckled and cannot fail early due to fatigue or pulsation. The buckling point is accurately repeatable with no adverse buckling point variation. The standard buckling point is ±5% of set point with ±2% available with valve test certificates. Features include: • • • •

The pin is external from aggressive system fluid. Proximity detector may be used to detect buckling. Buckling pin can be changed in minutes. Buckling pin valve accepted by: ASME Section VIII Division I Code Case 2091-3 API RP-520 Part One, Section 2.4 The National Board of Boiler and Pressure Vessel Inspectors   

10.14.2 Burst Disc Also known as a rupture disc, bursting disc, or burst diaphragm, the disc is usually made out of metal and designed to rupture at a predetermined pressure. Once the disc has ruptured, it will not re-seal (Figure 10.24). Apart from their low cost and almost instantaneous (milliseconds) response, rupture discs provide a number of other advantages that are specific to a wide range of applications: • Protection of the upstream side of a pressure relief valve against corrosion by the system fluid. • Protection of the upstream side of a pressure relief valve against plugging or clogging by viscous fluids or polymerization p­ roducts. • In place of a pressure relief valve if the protected system can tolerate process interruptions or loss of fluids in case the disc ruptures. • In place of a pressure relief valve if extremely fast response is desirable.

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Figure 10.24.  Also known as a rupture disc, bursting disc, or burst diaphragm, the disc is designed to rupture at a predetermined pressure and, once ruptured, will not re-seal (courtesy Oseco).

• As a secondary pressure-relieving device when the difference between the operating pressure and the rupture pressure is large, depending on the type of rupture disc selected. • To protect the downstream side of a pressure relief valve against downstream corrosion from headers or against atmospheric ­corrosion. • To minimize process/product leakage and reduce fugitive ­emissions. Bursting discs may be forward- or reverse-acting, depending on whether the pressure forces are acting on the concave (forward-acting) or convex (reverse-acting) faces (Figure 10.25). The operating pressure of forward-acting discs is usually limited to 65–85% of the disc’s predetermined bursting pressure. Reverse-acting discs are able to support pressures much closer to their rated burst pressure, typically up to 90% of the disc design bursting pressure. Rupture discs are usually held in place by a rupture disc holder (Figure 10.26), although some discs are designed to be installed between standard flanges without holders. Rupture discs are available in several configurations that include: • Conventional rupture disc (pre-bulged) • Scored tension-loaded rupture disc

Safety Relief Valves   •  215

Forward acting

Reverse acting

Figure 10.25.  Bursting discs may be forward- or reverse-acting.

Rupture disc Disc tab FLOW DIRECTION

Pressure

Figure 10.26.  Typical rupture disc holders (courtesy Oseco).

• Composite rupture disc (low pressure) • Graphite rupture discs (corrosion-resistant)

10.15 Conventional Rupture Disc A conventional domed rupture disc (Figure 10.27) is a pre-bulged solid metal disc designed to burst when it is overpressured on the concave side. They generally provide satisfactory service life when operating conditions are 70% or less of the rated burst pressure of the disc. Special designs are available for back-pressures exceeding normal atmospheric (15 psi). Conventional domed rupture discs will fragment on bursting.

10.16 Scored Tension-Loaded Rupture Disc Designed to open along scored lines, scored tension-loaded rupture discs (Figure 10.28) allow a closer ratio (generally 85%) of system operating pressure to disc burst pressure.

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Figure 10.27.  Conventional domed rupture discs are prebulged solid metal discs designed to burst when operating conditions are 70% or less of the rated burst pressure.

Figure 10.28.  Scored tension-loaded rupture discs allow a closer ratio (generally 85%) of system operating pressure to disc burst pressure.

10.17 Composite Rupture Disc Composite rupture disks (Figure 10.29) are designed for applications where lower rupture pressures are required than can be achieved with standard rupture discs. Typically, a composite bursting disc comprises a slotted metal top section together with a plastic membrane located on the concave or pressure side. The top section is the pressure zone and controls the bursting rating of the disc. Since the top section has open slots, the membrane isolates it from process media and prevents leakage. Composite rupture disks provide better corrosion resistance and are often used as corrosion barriers.

Safety Relief Valves   •  217

Figure 10.29.  Composite rupture disc (courtesy Continental Disc Corp.).

10.18 Graphite Rupture Disc A graphite rupture disc is manufactured from graphite impregnated with a binder material and designed to burst by bending or shearing (Figure 10.30). Graphite rupture discs are resistant to most acids, alkalis, and organic solvents. Operation to 70% of the rated burst pressure is generally permissible. A support may be required for discs that are rated 15 psi or less and for conditions of higher back-pressure.

Figure 10.30.  Graphite rupture disc manufactured from graphite impregnated with a binder material and designed to burst by bending or shearing (courtesy Svi Carbon Private Limited).

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Graphite rupture discs fragment upon rupture, and thus, provisions for capturing fragments may be required in certain applications. The service life of pre-bulged metal rupture discs under normal operating conditions is usually one year. They are subject to relatively rapid creep stress failure, especially at high operating temperatures. If not replaced periodically, they may fail without warning at normal operating pressures. Higher operating pressures (up to 90% of the disc design bursting pressure) are possible with the reverse-buckling disc. Less fatigue due to pulsating and cyclic operating pressure results in a longer service life than would be expected if the disc were installed with the pressure acting against the concave side. Most reverse-buckling discs should not be used in liquid full service. However, if an assured gas pocket rests against the disc and the disc manufacturer is consulted, liquid service can be considered. With finite lives, these discs should be replaced periodically. Consult the manufacturer for recommended replacement times.

10.19 Burst Disc Applications and Installation Practices Impervious graphite rupture discs offer nearly the same advantages and disadvantages as the reverse-buckling, metal type. However, with impervious graphite rupture discs, the piping arrangement may be more complicated, and uneven flange bolt loads or thermal strains in the piping may crack the disc. Rupture discs that tend to fragment, such as conventional and graphite discs, are typically not installed beneath pressure relief valves unless a means of protecting the pressure relief valve inlet from the fragments is provided. Caution: When rupture discs are removed for inspection or when an accompanying relief valve is serviced, the discs can easily be damaged and can fail prematurely if reused. Replacement of discs at every maintenance interval will minimize the chance of damage and premature failure. The proper receipt, storage, handling, and installation of a rupture disc are critical to its successful performance. The manufacturer’s installation instructions must be adhered to, especially those concerning limits on bolt torque. Some rupture discs using knife blades to open have failed to open properly. Consultation with the manufacturer concerning proper installation and maintenance of these kinds of rupture discs may be beneficial.

Safety Relief Valves   •  219

Tell-tale pressure indicator

Rupture disk

Figure 10.31.  Bursting disc installed on a safety valve.

A pressure gauge, a try cock, a free vent, or a suitable tell-tale indicator must be inserted between a rupture disc device installed at the inlet of a pressure relief valve and the valve (Figure 10.31), permitting the detection of disc rupture or leakage. Since rupture discs are designed to burst at a specified differential pressure, pressure build-up on the downstream side of the disc may inhibit the disc’s ability to provide overpressure protection. A rupture disc device must have full pipe area and must not fragment into the pressure relief valve inlet after the disc bursts. When a rupture disc device is used with a pressure relief valve, consult the ASME Code for the capacity reduction and installation details.

10.20 Performance Tolerance A simple way of understanding and comparing rupture disc specifications is to recognize the definition given by AS1358-1989 for Performance Tolerance (section 1, definition 1.2.8, as follows): ‘Performance tolerance—a range of pressure in positive and negative quantities or percentages which include both manufacturing range and bursting tolerance at a coincident temperature, which is applied directly to the specified bursting pressure.’ A rupture disc is usually specified using an min–max range of pressure at a specified temperature. When the min–max range is given by the supplier, the other specific details necessary to specify are:

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• The nominal burst pressure • The manufacturing range • The burst tolerance In some cases, suppliers try to confuse the issue by stating the min– max, but applying it to the manufacturing range only. This should be asked of the supplier to ensure their complete understanding of the true min– max as the true min–max incorporates not only the manufacturing range, but also all burst tolerances (Figure 10.32). The inclusion of manufacturing ranges and tolerances in the performance tolerance means that the batch of discs being ordered today and future batches will not burst outside this range at the at the customer’s specified coincident temperature. The testing carried out in the factory determines the actual burst pressure of the batch. If the customer nominates ASME VIII certification, two tests are made in an oven at the customer’s coincident burst pressure. The average of these tests must lie in the manufacturing range and is stamped on the disc tab in accordance with ASME VIII. If ASME certification is needed, then the stamping on the disc tab cannot vary. If stamping is needed to be min/max, users can specify the same stringent testing as the ASME code by stating that the testing shall include at least two tests at the elevated, coincident temperature that fall in the manufacturing range, e.g., they can then specify in accordance with ISO 6718. Essentially, if a user asks for a tighter min/max, he or she is asking for a tighter manufacturing range. Manufacturing ranges are specified in the manufacturers’ literature. A zero-manufacturing range is the tightest, meaning the average of the burst tests in the factory must equal the nominal burst pressure at the coincident temperature (Figure 10.33). The burst tolerance is always ±5% in accordance with all the rupture disc codes for stamped burst pressure equal to or greater than 276 kPa (g) Min Performance tolerance

+Z burst tolerance

Manufacturing range

Nominal burst pressure (High end of manufacturing range) Low end of manufacturing range

−Z burst tolerance Max

Figure 10.32.  Performance tolerances.

Safety Relief Valves   •  221

Performance tolerance for zero manufacture range rupture disk

Max +Z burst tolerance Nominal burst pressure = Stamped burst pressure −Z burst tolerance Min

Figure 10.33.  A zero-manufacturing range is the tightest, meaning the average of the burst tests in the factory must equal the nominal burst pressure at the coincident temperature.

at 22ºC. The burst tolerance varies according to the particular disc design for stamped burst pressures below 276 kPa (g) at 22ºC. The burst tolerance refers to the accuracy of each disc in the batch received. Essentially, once it has been confirmed what manufacturing range (ask the manufacturer to specify low end and high end) and what burst tolerance apply to the high end and low end of the manufacturing range, then the performance tolerance is worked out. If the supplier cannot provide this detail, then they should confirm that all manufacturing ranges and tolerances are included in the min/max they have given.

10.21 Maximum Operating Pressure Any rupture disc can be operated to any desired pressure. However, at what pressure should users operate for reliable performance, so that the life of the disc is not reduced? From more than 65 years’ experience in the field, BS&B (a leading manufacturer) has established the following operating pressure to burst pressure ratios (expressed as %) (Table 10.1).

10.22 Cyclic/Pulsating Duties The next question is: which discs can be operated reliably in cyclic/pulsating duties? The answer is S90, JRS, RLS, MRB, and ECR. All others are tension-loaded discs that will have a limited cycle, life but are very good in static conditions.

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Table 10.1.  Typical operating pressure to burst pressure ratios (courtesy BS&B)

BS&B types S90, JRS, CSR, RLS, MRB, ECR GFN XN, LCN, DV EXP/DV

AV, AVV BV

MBV

Operate from full vacuum to Y% of the Description stamped BP Single-section reverse-buckling 90% disc Tension-loaded disc that is 85% scored after disc crowned Composite tension-loaded disc 80% Tension-loaded, domed, 80% of nominal composite disc with vacuum support Flat composite disc with 60% of gaskets minimum Tension-loaded, solid metal 70% pre-bulged disc with vacuum support Integral disc and holder of 80% high-grade impregnated graphite with vacuum support

Hence, the operating ratio can only be qualified depending on the type of service that the disc is used in. This optimum value must, therefore, be seen with caution. On the specification sheet, the next step is to specify the maximum positive operating pressure that the disc should see, at the coincident burst temperature to ensure maximum service life. To specify this, for the worst case condition, users must establish first whether the low end of the manufacturing range will be above or below 276 kPa (g) (Figure 10.34).

10.23 Case A: 276 kPa (g) or higher In this case, users may operate to Y% of the stamped burst pressure (at worse case Y% of the minimum of the manufacturing range).

Safety Relief Valves   •  223

Min Performance tolerance

+Z burst tolerance Nominal burst pressure Manufacturing range −Z burst tolerance Max Y% × Min of low end of manufacturing range is < 276 kPa(g)

Y% × Min of low end of manufacturing range is > 276 kPa(g)

Maximum operating pressure

Maximum operating pressure Figure 10.34.  Determining the upper and lower maximum operating ranges according to whether the pressure lies above or below 276 kPa (g).

10.24 Case B: Lower than 276 kPa (g) In this case, you may operate to Y% of the min, and hence, the burst tolerance must be deducted from the stamped burst pressure (at worse case, the burst tolerance is deducted from the low end of the manufacturing range and the min is calculated. Then Y% is applied to the min). The performance of rupture discs at temperatures other than the coincident disc temperature cannot be guaranteed unless the user is prepared to pay for extra testing. Estimates can be given for various disc materials, although these should not be taken for granted.

10.25 Standards Commensurate with that of the different definitions described, there are also a wide number of standards. Furthermore, national standards define many varying types of safety valve. The performance of different types of safety valve, set out by the various standards, is summarized in Table 10.2.

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Table 10.2.  Safety valve performance summary Standard

Fluid Steam

A.D. Merkblatt A2 Air or gas Liquid I ASME

BS 6759

VIII

Standard 10% full lift 5% 10% Steam Steam Air or gas Liquid

Part 1

Steam

Part 2 Part 3

Air or gas Liquid

Overpressure Standard 10% full lift 5% 10%

20% 3% 10% 10% 10% (see note 3) Standard 10% full lift 5% 10% 10–25%

Blowdown 10%

2–6% 7% 7%

10% 10% 2.5–20%

APPENDIX A

J–T Valve The J–T (Joule–Thomson) valve is frequently used in the p­ etrochemical industry to liquefy gases and is essentially an expansion valve, often referred to as a ‘choke’ or ‘choke valve.’ The Joule–Thomson effect describes how, when a gas is forced through a restriction and then expands, the average distance between the molecules increases. Because of intermolecular attractive forces, the gas, thus, undergoes an increase in the potential energy, and assuming no heat is transferred during this process, the total energy of the gas remains the same. The increase in potential energy is matched by a corresponding decrease in the kinetic energy, and therefore in temperature. There is also a reverse process (limited to hydrogen, helium, and neon gases) in which the temperature actually increases as the gas expands. Called the ‘reverse Joule-Thompson effect,’ as the gas expands and the intermolecular distances increase, the number of collisions will fall and cause a decrease in the average potential energy. Again, since the total energy of the gas remains the same, there will be an increase in the kinetic energy, and therefore in temperature. It should also be noted that, at very high pressures, in the order of 600 bar, many naturally occurring hydrocarbon gases are subject to the ‘reverse Joule-Thompson effect’ and will heat on expansion.

J–T Valve Construction The term ‘J-T valve’ refers to the application, rather than the valve itself. That being said, there are certain physical constraints of any valve used in cryogenic applications.

226  •   J–T Valve

Although use may be made of any throttling valve, angle needle valves have proved popular in the oil industry, although straight cage/ globe valves are also frequently used. Probably, the first requirement is to select the right materials of construction. At cryogenic temperatures, valve body materials would exclude carbon steel and include austenitic stainless steel, bronze, and Monel. A second area of concern in cryogenic applications is packing. While PTFE is restricted to around –40°C, graphite packing generally caters for temperatures down to –200°C. An area of more general concern, however, is that moisture from the atmosphere condenses on the colder surfaces and forms a layer of frost, and even ice, on the bonnet and stem areas of the valve. As the stem is stroked, frost may be drawn into the packing area and cause damage. This problem may be overcome through the use of an extended valve stem and bonnet that places the packing box area far enough away from the cold area of the valve to prevent freeze-up of the packing and minimize damage. The bonnet would generally be made of austenitic stainless steel to minimize heat transfer. Some form of insulating ‘cold box’ surrounding the valve and its piping, to minimize heat exchange with the environment, is also helpful in combating the effects of both moisture build-up and noise. Another basic problem associated with these J–T valves is as a result of water or other contaminants freezing out of the gas. The formation of both water ice and CO2 dry ice can plug the valve. This may be overcome by dehydrating the gas well below the recommended dewpoint.

APPENDIX B

Basic Acoustics The term ‘acoustics’ applies to a branch of physics concerned with the properties of sound. And, ‘sound’ implies that it is something we hear― that it is detected by the ear. All sound originates from a vibrating body and is generated when we displace the normal random motion of air molecules. If you pluck the string of a musical instrument, the air molecules are displaced to produce alternate rarefaction (expansion) and compression.

Pitch (Frequency) One of the most important properties of sound is its pitch or tone. As we have seen, sound is basically a vibration of molecules, and it is the rate of vibration that determines the pitch (how high or low) of the resultant note. The normal method of illustrating the alternate rarefaction and compression of sound is in the form of a waveform, called a sine wave (Figure B.1). The number of vibrations or cycles or repeats of an event each second is called the frequency and is represented by the unit hertz (Hz), named after the German physicist Heinrich Hertz. Thus, 100 cycles or vibrations per second would be designated 100 Hz. It is also normally customary to abbreviate 10,000 Hz to 10 kHz, where the k signifies kilo or 1,000. The range of vibration audible to human beings ranges from 20 to 20,000 Hz (20 kHz), encompassing all musical instruments plus the human voice. The range of 20 to 20 kHz normally applies only to a young healthy person of about 25 years. As people age, so their high frequency response falls off rapidly until at the age of 60 years, few are able to detect frequencies above 10 kHz.

Amplitude

228  •   Basic Acoustics

1 cycle

Figure B.1.  Graphic representation of a sound wave showing frequency and amplitude.

Timbre If a Middle C is played on a piano, on a violin, and on a trumpet, we have learned that it will produce the same tone of 256 Hz in each case. Nonetheless, the tone produced by the piano is easily distinguishable from that played on the violin, and in turn from that played on the trumpet. What distinguishes the three instruments is their timbre. If a violin string is plucked in the middle of its length, it will produce a tone that is determined by the tension and the mass of the string. By adjusting the tension, this frequency can be adjusted to, for example, Middle C (256 Hz). This is called its fundamental frequency. If the string is now lightly touched in the middle and plucked at a distance that is a quarter of its length, it will vibrate in two segments and give rise to a frequency that is twice that of the original or fundamental frequency. And, by lightly touching a point one-third of its length and plucking half-way between that point and the end, the string will vibrate in three segments and produce a frequency that is three times higher than the original or fundamental. These higher frequencies are all a characteristic of the same length of string and are called harmonics or overtones of the fundamental. When these harmonics are added to the fundamental, they give rise to a waveform that, although containing the tone of the note, also makes an individual sound.

Basic Acoustics   •  229 2nd harmonic

Fundamental

(a)

(b)

Fundamental

3rd harmonic

Figure B.2.  Addition of harmonics to the fundamental: (a) second harmonic; (b) third harmonic.

Figure B.2(a), for example, illustrates a second harmonic added to the fundamental, while Figure B.2(b), illustrates a third harmonic added to the fundamental.

Velocity The third property of sound that we need to look at is its velocity through the medium in which it is traveling. The velocity of sound in gases is given by the equation:

C =

g ⋅ R ⋅ ( 273 + T ) M

where: C = velocity of ultrasonic wave (m/s) R = universal gas constant 8,314.3 (J/K.mol) T = temperature (°C) M = molecular weight (kg/K.mol) γ = adiabatic component

230  •   Basic Acoustics

In air, the most common medium used for the propagation of sound, the velocity, may be calculated by: 20.048

( 273.15 + T)

Therefore, at 20°C: C = 20.048 293.15 ≈ 343 m / s This means that the sound heard at a distance of 343 m from its source reaches the ear after 1 second. Put another way, the sound from a lightning flash takes about 3 seconds to travel each kilometre distance. In most media, other than air, sound travels faster. In water, for ­example, sound travels four times faster than in air, and in steel, it travels some 14 times faster.

Wavelength Assume a sound wave traveling in air at a frequency of 1 kHz. Thus, in one second, there will be 1,000 peaks during which the sound will have traveled 343 metres. So, on one full cycle (peak to peak), the wave will have travelled only 1/1,000th the distance i.e., 0.343 metre. This distance is called the wavelength (denoted by the symbol λ) and is obtained by dividing the velocity by the frequency. Thus:

l=

V f

The wavelengths of several frequencies traveling in air are shown in Table B.1. Table B.1.  The wavelengths of several frequencies travelling in air Frequency 500 Hz 1 kHz 2 kHz 4 kHz 8 kHz 16 kHz

Wavelength 686 mm 343 mm 171.5 mm 85.8 mm 42.88 mm 21.44 mm

Basic Acoustics   •  231

Intensity If a violin string is plucked with varying degrees of force, the pitch of the note remains the same, but its intensity (or loudness) increases. Referring to the waveform of Figure B.1, the intensity can be represented by the height or amplitude. The greater force used to pluck the string, the louder the sound and the higher the amplitude.

Logarithmic Characteristic of Ear A remarkable ability of the human ear is its ability to cope with a huge range of sound intensities Ñ ranging from the fall of a pin through to the scream of a modern jet engine. This ability of the ear to cope with this range is due to its non-linear characteristic. Figure B.3 shows two identical loudspeakers, each connected across an amplifier. The amplifiers are driven by a 1 kHz tone generator that is switched so that it is connected, alternately, to each. If the same power is supplied to each speaker, e.g., 100 mW, both their notes will be of equal intensity, and a listener will perceive them to be the same. Initially, as the power to one of the speakers is gradually increased, the listener, listening to each of them one at a time, will perceive no difference in intensity. Only when one loudspeaker receives 26% more power will it in fact sound louder. At this point, 126 mW is being fed to one loudspeaker and 100 mW to the other.

Figure B.3  Two identical loudspeakers connected across an amplifier.

232  •   Basic Acoustics

If the balance is now restored, by bringing both loudspeakers up to 126 mW, the intensities will again be equal. Once more, if the power to the one loudspeaker is increased, no perceptible difference will be noticed until it receives 26% more power, e.g., 26% of 126 mW = 32 mW, thus bringing the higher loudspeaker output to: l26 + 32 = 158 mW. Once again, if the balance is restored by bringing the other loudspeaker up to 158 mW, the intensities will again be equal. Now, again, the intensity of one speaker will need to be 26% higher, i.e., 26% of 159 mW = 41 mW, before the listener perceives a difference. This brings the higher loudspeaker output to: 159 + 41 = 200 mW. If this procedure is repeated in 10 stages, it will be found that the intensity has also increased by 10 times to 1,000 mW. The perceived increase in the sound intensity is, thus, obtained by raising the power level in a given ratio—not by adding specific amounts. The important point to note is that this 10-fold increase has not been reached by 10 equal linear steps, but by 10 non-linear steps, the first being quite small and the last quite large (Figure B.4). This response is called logarithmic. 1000 900 800 700

mW

600 500 400 300 200 100

1

2

3

4

5

6

7

Step number

Figure B.4.  Logarithmic response of the human ear.

8

9

10

Basic Acoustics   •  233 10 dB

1000 mW

9 dB 794 mW

8 dB

mW

630 mW

7 dB 501 mW 6 dB

398 mW

5 dB 316 mW 4 dB 3 dB 200 mW 0 dB

126 mW

251 mW 158 mW 100 mW

Figure B.5.  The 10-fold power ratio increase is designated a Bel with each power increment of 26% being one-tenth of a Bel—called a decibel (dB).

As illustrated in Figure B.5, the 10-fold power ratio increase is designated a Bel with each power increment of 26% being one-tenth of a Bel, called a decibel (dB). It must be appreciated that the dB is only a ratio, and that the ear hears the same small difference between 1 W and 2 W, as between 100 W and 200 W.

Definition of Decibel The mathematical definition of the decibel is: 10 log10

P2 P1

where: P2 = output power P1 = input power In practice, fortunately, it is not necessary to make use of this equation, as tables are given for any ratio that is required (Table B.2).

234  •   Basic Acoustics

Table B.2.  Gain and attenuation ratios expressed in dBs

Power ratios (A) 1 1.26 1.585 2 2.5 3.162 3.98 5 6.31 7.943 10 12.6 15.9 20 25.1 31.6 39.8 50.1 63.1 79.4 100 1 000 104 105 106 107 108 109 1010 1011 1012 1013 1014

Gain Voltage, current, and pressure ratios (B) 1 1.122 1.26 1.413 1.585 1.778 2 2.24 2.5 2.82 3.162 3.55 3.98 4.47 5 5.62 6.31 7.08 7 943 8 91 10 31.6 100 316 103 3.16 × 103 104 3.16 × 104 105 3.16 × 105 106 3.16 × 106 107

Attenuation Voltage, current, dB Power and pressure +ratios ratios ‹fi (C) (D) 0 1 1 1 0.79 0.89 2 0.63 0.79 3 0.5 0.7 4 0.4 0.63 5 0.32 0.56 6 0.25 0.5 7 0.2 0.45 8 0.16 0.4 9 0.13 0.355 10 0.1 0.32 11 0.08 0.28 12 0.063 0.25 13 0.05 0.224 14 0.04 0.2 15 0.032 0.18 16 0.025 0.16 17 0.02 0.14 18 0.016 0.126 19 0.013 0.112 20 0.01 0.1 30 0.001 0.0316 −4 40 10 10−2 −5 50 10 3.16 × 10−3 60 10−6 10−3 −7 70 10 3.16 × 10−4 80 10−8 10−4 −9 90 10 3.16 × 10−5 100 10−10 10−5 −11 110 10 3.16 × 10−6 120 10−12 10−6 −13 130 10 3.16 × 10−7 140 10−14 10−7

Basic Acoustics   •  235

Voltage Ratios So far, we have discussed power ratios, and thus an amplifier having a 3 dB gain will double the power. The question arises, what would happen to the voltage in the same amplifier? In this circuit (Figure B.6), the power developed across the resistor is given by: P=

V 2 l00 = = 10 W 10 R

How much would we need to increase the voltage if we wanted to double the power to 20 W (i.e., increase of 3 dB)? Since: P =

V2 R

V2 = P ⋅ R V=

P⋅R =

20 ⋅10 = 14.4V

As illustrated in Figure B.7, doubling the power is achieved by only a 1.414 increase in voltage. Using a similar reasoning, it can be shown that: 3 dB = 2 × power but only 1.414 × voltage

10 V

10 Ω

10 W

Figure B.6.  Simple circuit showing power developed by a resistor.

236  •   Basic Acoustics

14.4 V 20 W 10 V

10 Ω 10 W

Figure B.7.  Doubling the power is achieved by only a 1.414 increase in voltage.

6 dB = 4 × power but only 2 × voltage 10 dB = 10 × power but only 3.16 × voltage 20 dB = 100 × power but only 10 × voltage

Absolute Levels As stressed earlier, the decibel is only a ratio. Nonetheless, it can be used to express absolute values if a reference is given. If the reference chosen was 1 W, then 3 dB would be 2 W and 6 dB would be 3.98 W and so on. In fact, one of the common references used is the milliwatt = 10−3 W and is designated dBm. Thus: 20 dBm = 100 ¥ 1 mW = 100 mW. Another reference that is used is in the measurement of sound ­pressure  levels.

Sound Pressure Level (SPL) Previously we have seen that sound is a movement or vibration of ­molecules giving rise to increases or decreases of pressure. How do we measure air pressure? Pressure is defined as the force per unit area, and for many, the definition of pounds per square inch (lbs/in2 or psi), still prevalent in the

Basic Acoustics   •  237

United Kingdom and United States, seems very much more descriptive of pressure rather than, for example, the Pascal. However, from physics, we learned that another useful definition is the bar: 1 atmosphere = 14.7 lbs/in2 = 29.92 inches of mercury (Hg)

= 760 mm Hg



= 1 bar.

Now clearly, 1 bar is far too large a unit in which to express changes in sound pressure, and therefore the microbar (µbar) is often used where 1 µbar is one millionth of a bar: 1 µbar = 1/1,000,000 bar

= 10−6 bar.

With the introduction of the metric system, the microbar was replaced by the Pascal where the Pascal (Pa) is one-hundred-thousandth of a bar: 1 Pa = 1/100,000 bar

= 10−5 bar.

From this, it follows that: 1 Pa = 10 µbar

Standards The absolute standard that has been chosen for measuring the sound pressure level (SPL) is called the threshold of hearing and is the average sound pressure that is barely perceptible to a young human, with undamaged hearing, at 1 kHz. This level is generally described as 0.0002 µbar or 0 dB (SPL). In reality (as shown in Figure B.8), the actual threshold is about 4 dB (0.003 µbars) higher at 1 kHz and is very frequency-dependent such that the lower the frequency the more energy required to make it audible. At 50 Hz, for example, the average person cannot hear SPLs below 40 dB (0.02 mbars) As the SPL is increased, we reach a point just short of being painful to the ear, called the threshold of pain which, using the 1 kHz datum, ­corresponds to 200 µbars. Since the absolute reference of 0 dB is 0.0002 µbar, the value for the threshold of pain is 200/0.0002 = 106 = 1 million times greater. This may be expressed in dBs by referring to Table B.3, column B where: 106 = 120 dB (SPL)

Sound pressure level (SPL dB)

238  •   Basic Acoustics 130 120 110 100

Threshold of pain

90 80 70 60 50 40 30 20

Threshold of hearing

10 0 –10 10 Hz

100 Hz

1 kHz

10 kHz

100 kHz

Frequency (Hz)

Figure B.8.  Threshold of hearing (lower) and of pain (upper).

Table B.3.  Sound intensities of various sources Source Jet plane at 30 m Threshold of pain Indoor rock concert Siren at 30 m Male voice shouting Motor car interior moving at 90 k/h Busy street traffic Average male voice at 1 metre Average car Quiet radio Ticking clock Whisper Rustle of leaves Threshold of hearing

Other important levels are: 1 mbar = 74 dB (SPL) 1 Pa = 10 mbar = 94 dB (SPL)

Intensity level dB (SPL) 140 120 120 100 87 75 70 67 60 40 30 20 10 0

Basic Acoustics   •  239

As discussed, the energy level at the threshold of hearing varies with frequency. Thus, it would require 54 dB (SPL) at 30 Hz to produce the same ‘loudness’ (as perceived by the listener) as 0 dB (SPL) at 1 kHz. The contour for the threshold of hearing represents only the bottom limit of a series of ‘equal loudness contours’ called phons (Figure B.9). Two factors are very much evident from the equal loudness contours of Figure B.9. Firstly, very much more energy is needed to produce a bass note of a given loudness, when compared with a 2–3 kHz note—an important consideration in any music system. The second point is that, if a broad spectrum noise level of, say, 20 dB (SPL) is introduced, the listener will have a certain sound ‘image’ according to the equal loudness contour corresponding to 20 dB. If the noise level is now raised, another sound image is received, of the same noise, due to the mechanism of the changing curves of equal loudness. In other words, all frequencies are present in the signal, but depending on the level, will be heard in different relationships. In order to reproduce this subjective perception characteristic of the ear, a sound level meter should incorporate automatic switching filters that represent the phon curves at every level. Because this would increase the cost of such instruments to an unaffordable level, standard weighting curves have been determined (Figure B.10):

Sound pressure level (SPL dB)

130 120

Threshold of pain

110 100 90 80 70 60 50 40 30 20 10

Threshold of hearing

0 –10 10 Hz

100 Hz

1 kHz Frequency (Hz)

Figure B.9.  Equal loudness contours.

10 kHz

100 kHz

240  •   Basic Acoustics 20 10

Gain (dB)

0 –10 A-weighting B-weighting

–20 –30

C-weighting

–40 –50 10 Hz

100 Hz

1 kHz

10 kHz

100 kHz

Frequency (Hz)

Figure B.10.  A- B-, and C-weighted responses required for measuring sound pressure levels.

A-curve Designated dBA, this weighting was originally intended only for the measurement of low-level sounds (around 40 phon). B-curve This curve is applied for levels between 40 dB and 70 dB. C-curve Used at levels higher than 70 dB. Note that all three curves cross at 1 kHz. A-weighting is now mandated for the measurement of environmental noise and industrial noise. In reality, it is badly suited for these purposes, since it tends to moderate the effects of low frequency noise.

Appendix C

Block and Bleed The primary function of a double block-and-bleed system is to isolate or block the flow of the upstream process medium from reaching the downstream equipment while carrying out maintenance, repair, or component replacement. As illustrated in Figure C.1, it typically comprises two block valves (Valves 1 and 2) and a bleed valve (Valve 3) that vents to a relief or safe disposal location. When delivering the process medium to the downstream equipment, the valves are set with the isolation valves 1 and 2 open and the bleed valve 3 closed. When isolating the downstream equipment from the process fluid, the valves are set with isolation valves 1 and 2 closed and bleed valve 3 open. By monitoring the outlet of the bleed, users may determine whether the downstream system is, in fact, properly isolated and whether there is any leakage past either of the two block valves. Usually, a double block-and-bleed would comprise two separate block valves and a separate bleed valve assembled on a tee. Block valves can take the form of virtually any high-integrity shut-off valve (e.g., ball valve, expanding gate valve, or plug valve), while the bleed is usually a ball valve, or in many cases, a cap or plug. Although such multi-block valve systems work effectively, they can be expensive to install and maintain, especially when dealing with largesized valves or working with automated systems. Consideration must be given not only to the capital cost of the two block valves; bleed valve; actuation, four or three valves; control system, type T; flange bolting; and the flange gaskets, but also the design and installation costs. Consequently, several manufacturers have designed single manifold systems, such as that shown in Figure C.2, which illustrates the typical construction of a single double block-and-bleed manifold

242  •   Block and Bleed Process medium in

OPEN

Valve 1

Valve 2

Valve 3

(a)

Process medium in

(b)

OPEN

CLOSED

CLOSED

CLOSED

Valve 1

Valve 2

Valve 3

OPEN

Process vessel

Process vessel

To relief system or safe disposal

Figure C.1.  (a) Under normal operation, the valves are set with the isolation valves 1 and 2 open and the bleed valve 3 closed. (b) When isolating the downstream equipment, the valves are set with isolation valves 1 and 2 closed and bleed valve 3 open.

Figure C.2.  Typical construction of a single double blockand-bleed valve (courtesy Habonim).

Block and Bleed   •  243

Valve body

Process fluid

Figure C.3.  The bleed often takes the form of a cap or plug.

system comprising two large balls acting as blocks (both shown closed) and a small ball serving as the bleed (ball is shown in the open position). In addition, specially designed trunnion-mounted ball valves, equipped with a valve body bleed between the seats, provide a satisfactory substitute for separate individual double block-and-bleed valves (see Chapter 4. ‘Valve Construction,’ Section 4.9). The bleed often takes the form of a cap or plug as illustrated in Figure C.3.

APPENDIX D

Water Hammer Because liquid is essentially incompressible, any energy applied to it is transmitted instantly. If a moving column of liquid is slowed down ­suddenly by, for example, a quick-closing valve, the sudden change in liquid velocity in the delivery line creates a pressure wave (Figures D.1(a) and (b)). Despite the frequent assumption that liquids are incompressible, in reality, most substances diminish in volume when exposed to a uniform externally applied pressure. The Bulk Modulus describes the compressibility of a fluid as the ratio of the very small decrease in volume resulting from an applied external pressure. A large bulk modulus indicates a relative incompressible fluid. Assuming that the walls of the pipe are sufficiently thick for it to be approximated as rigid, the velocity of the pressure wave in a rigid pipe is given by: C=



K  ρ

(D.1)

where: C = velocity of pressure wave (m/s) K = bulk modulus (GPa) ρ = density (kg/m3) Thus, for example, given that water has a density (ρ) of 1,000 kg/m³ and a bulk modulus of 2.2 GPa, the velocity is:

C=

2.2 ⋅109 = 1483m/s  1000

(D.2)

Some typical sonic velocities in various liquids are shown in Table D.1.

246  •   Water Hammer Large diameter riser Open valve (a) Branch

Flow

Normal flow

Valve closed (b) Shock Quick closure

Figure D.1.  If a moving column of liquid (a) is slowed down suddenly by, for example, a quick-closing valve, the sudden change in liquid velocity in the delivery line creates a pressure wave (b).

Table D.1.  Some typical velocities in various liquids Sound velocity at 25°C (m/s) (ft/s) 1,324 4,344 1,250 4,101 1,483 4,862 1,347 4,420 1,401 4,598 1,441 4,729 1,480 4,856

Liquid Kerosene (paraffin) Gasoline (petrol) Water Crude oil: Light Crude oil: Medium Crude oil: Heavy Crude oil: Extra heavy

In practice, the pressure rise may be sufficient to deform the pipe, increasing its cross-section. And, since the pipe thus absorbs strain energy, the velocity of the pressure wave is reduced:

C=

Ke  ρ

(D.3)

where: Ke = the effective bulk modulus, given by:

1 1 D  = + Ke K E T

(D.4)

Water Hammer   •  247

where: K = bulk modulus of the liquid (GPa) D = internal diameter of the pipe (mm) E = bulk modulus of the pipe (GPa) T = thickness of the pipe (mm) Again, assuming water (density (ρ) 1,000 kg/m³ and a bulk modulus of 2.2 GPa) flowing in a pipe having a bulk modulus (E) of 210 GPa, an internal diameter of 200 mm, a wall thickness of 5 mm: From equation (D.4): 1 1 200 1 (D.5) = + = K e 2.2 ⋅109 210 ⋅109 ⋅ 5 6.45 ⋅1010



Therefore, the effective bulk modulus (Ke) is 1.55*109, and substituting:

C=

1.55 ⋅109 = 1244m/s (D.6) 1000

This corresponds to a substantial 16% reduction in the sonic velocity. In practice, the pressure wave travels back up the line at between 1,000 and 1,300 m/s, to the end of the pipe where it will reverse direction and travel back toward the valve (Figures D.2 (a) and (b)).

(a) Pressure wave enlarges pipe

(b) Reflected pressure wave

Figure D.2.  The pressure wave travels back up the line (a) at between 1,000 and 1,300 m/s, to the end of the pipe where it will reverse direction and travel back toward the valve (b).

248  •   Water Hammer

Pressure wave reaches valve

Figure D.3.  Depending on the valve size and system conditions, a valve closing in 1.5 s or less can produce a pressure spike five times the system working pressure.

Depending on the valve size and system conditions, a valve closing in 1.5 s or less can produce a pressure spike five times the system working pressure (Figure D.3), leading to blown diaphragms, seals, and gaskets and also catastrophic system component failure in transmitters, meters, and gauges. As intimated, the magnitude of the pressure spike in a given system is very much determined on the speed at which the valve is closed. Although there are many calculations available, a general rule of thumb is shown as follows:

P=

0.052 v L + PI t

(D.7)

where: P = increase in pressure (bar) v = flow velocity (m/s) t = valve closing time (s) L = upstream pipe length (m) PI = inlet pressure (bar) In FPS terms, the equation becomes:

P=

0.07 v L + PI  t

(D.8)

where: P = increase in pressure (psi) v = flow velocity (ft/s) t = valve closing time (s) L = upstream pipe length (ft) PI = inlet pressure (psi) To give you some idea of the magnitude of the spike, assume a solenoid valve having a closure time of approximately 40 to 50 ms, connected

Water Hammer   •  249

to a 15-m-long upstream pipe. The water flow is 3 m/s and the inlet delivery pressure is 4 bar. What is the amplitude of the pressure spike? From equation (D.7):

P=

0.052 3 15 + 4 = 62.5 bar  0.04

(D.9)

Water Hammer as a Result of Steam Condensate Water hammer also occurs as a result of condensate in steam pipes. Figures D.4 (a) to (d) show accumulated condensate in a portion of horizontal steam piping. As the steam flows over the condensate, it causes the surface of the water to ripple and trap some of the condensates in the pipe. At the same time, as a result of the Bernoulli effect (Figure D.4 (a)), a wave is drawn up that effectively seals the pipe, producing an isolated pocket of steam (Figure D.4 (b)). The collapsing steam void (Figure D.4 (c)) creates an implosion (Figure D.4 (d)) that produces a slug of condensate that is carried along by the steam flow and that can travel at the speed of the steam (up to 160 km/hr). The effect of this force, striking the first elbow in its path, is comparable to a hammer blow and the damage sustained can be quite substantial. Condensing steam

Heat loss Steam

(a)

Bernoulli effect draws up wave

Sub-cooled condensate Isolated steam pocket

Heat loss Steam

(b) Sub-cooled condensate

Wave seals pipe

Collapsing steam void 5 bar steam

(c) Implosion (d)

Rebounding wave

Figure D.4.  Hydraulic shock wave produced as a result of accumulated condensate in steam piping.

250  •   Water Hammer

A second type of water hammer that occurs in steam piping is actually cavitation. This is caused by a steam bubble forming or being pushed into a pipe completely filled with water. As the trapped steam bubble loses its latent heat, the bubble implodes, the wall of water comes back together, and the force created can be severe. This condition can crush float balls and destroy thermostatic elements in steam traps. This type of cavitation usually occurs in wet return lines or pump discharge piping.

Pulsations Of course, such shocks are not just produced by the closure of a valve. Other causes include: starting or stopping a pump, closure of an ESD device, and shut-off of a check valve. Pulsations are also often introduced through the use of ‘Oval’ gear positive displacement flow meters or ­reciprocating or peristaltic positive displacement pumps. The resultant acceleration and deceleration of the pumped fluid produces pressure spikes of greater than 10 times the steady state flow pressure.

Prevention and Mitigation The most obvious solution is, of course, prevention. Do not ever close a valve, do not ever trip or start a pump, do not have an emergency disconnect of a hose, etc. Clearly, wishful thinking! However, it is possible, in many cases, to close the valve under controlled conditions—increasing the valve closure time. In the example given previously, what would be the result of increasing the closure time to 1.5 seconds?

P=

0.052 3 15 + 4 = 5.56 bar  1.5

(D.10)

There are of course many ‘rule of thumb.’ One frequently used approximation is that the valve should not close faster than the acoustic round trip, which gives a ballpark figure that closure should be no faster than 30 s. Indeed, in U.S. waterways under the jurisdiction of the U.S. Coast Guard, a discharge valve is not allowed to close faster than 30 s when loading a tanker (Code of Federal Regulations 33CFR154—a USCG regulation). If we look at a typical example of a discharge from an SPM (single-­ point mooring) to the crude oil terminal through a 16 inch 5 km pipeline and assuming a sonic velocity of 3,300 ft/s, the acoustic round trip is just under 10 s—well below the accepted norm of 30 s.

Water Hammer   •  251

However, what sort of overpressure surge are we still likely to expect? Assume a flow velocity of 2 m/s (fairly slow) and a pipeline pressure of 6 bar, then:

P=

0.052 2 5000 + 6 = 23.3 bar (D.11) 30

With typical maximum allowable operating pressures (MAOPs) of the order of only 18 to 20 bar, this is certainly exceeding operational limits, even allowing for the fact that many codes allow for a 10% exceedance of the MAOP in the event of hydraulic surge. In very large pipelines, the use of motorised operated gate valves may have closure times of up to 4 or 5 min. Even so, on a 300-km-long pipeline, the overpressure surge is likely to exceed the MAOP. Since total prevention might well prove impossible to achieve, the answer must lie with mitigation through surge relief systems. In essence, there are three forms of relief devices available: pulsation dampeners, rupture discs, and surge relief valves. Pulsation Dampeners A pulsation dampener or surge suppressor is a hydro-pneumatic dampener comprising a pressure vessel containing a compressed gas, generally air or nitrogen, separated from the process liquid by a bladder or diaphragm. These devices are mainly used to minimize the pulsations resulting from a pump’s stroking action. During the discharge stroke, fluid pressure displaces the bladder and compresses the trapped gas (Figure D.5). During

Air/gas

Bladder

Liquid

Figure D.5.  A pulsation dampener or surge suppressor. During a surge, the fluid pressure displaces the bladder and compresses the trapped gas.

252  •   Water Hammer

the following cycle, the momentary interruption of fluid flow causes the compressed gas to expand, forcing the bladder or bellows to push the accumulated fluid back into the discharge line. An advantage of this type of system is that, it is ready for immediate reuse after a pressure surge has occurred. On the negative side, a single device can only relieve a small amount of fluid, and thus, on larger pipeline systems, a large bank of accumulators may be required. Nitrogen-Loaded Surge Relief Valves A typical gas-loaded axial flow style valve, from Daniel, is shown in Figure D.6, in which nitrogen gas is used to pressurize the valve piston to keep it in the closed position. The gas pressure less the 4 psi force of the valve spring is the effective set-point of the valve. The oil acts as a movable barrier between the gas and valve piston that eliminates the ­possibility of gas bypassing the piston. As the pipeline pressure increases, the combined force of the spring and nitrogen gas pressure is overcome and the valve opens (Figure D.7), with a response time typically under 100 ms.

Gas pressure Light oil Check valve Spring Piston

Flow

Figure D.6.  A Daniel gas-loaded axial flow style valve in which nitrogen gas is used to pressurize the valve piston to keep it in the closed position (courtesy Emerson).

Water Hammer   •  253 Gas pressure expelled Sight gauges Light oil Check valve Spring Piston

Flow

Figure D.7.  As the pipeline pressure increases, the combined force of the spring and nitrogen gas pressure is overcome and the valve opens (courtesy Emerson).

APPENDIX E

Stainless Steel Stainless steel is a family of corrosion-resistant steels containing chromium in which chromium forms a passive film of chromium oxide (Cr2O3) when exposed to oxygen. This phenomenon is called passivation and is seen in other metals, such as aluminum and titanium. The film layer is impervious to water and air and quickly reforms when the surface is scratched. This protects the metal beneath, preventing further surface corrosion. Since the layer only forms in the presence of oxygen, corrosion resistance can be adversely affected if the component is used in a non-oxygenated environment, e.g., underwater keel bolts buried in timber. Such passivation only occurs if the proportion of chromium is high enough and is normally achieved with addition of at least 13% (by weight) chromium. Progressively higher levels of corrosion resistance and strength is achieved by the addition of other alloying elements, each offering specific attributes in respect of strength and corrosion resistance.

Classification Issues The need to classify stainless steel has led to a fundamental problem of which method to use. Probably, the best known system derives from of the Society of Automobile Engineers (SAE), e.g., 316 Cr/Ni/Mo 17/12/2. This is interpreted as stainless steel containing the proportions of 17% chromium, 12% nickel, and 2% molybdenum. However, the waters are somewhat muddied by a variety of international and country-based systems that include EN (European Norm) and UNS (Unified Numbering System). For example, SAE 304 Cr/Ni 18/10 stainless steel is EN 1.4301, which is UNS S30400.

256  •  Stainless Steel

Stainless steels may also be graded into five basic families or phases determined by their crystalline structure: the stable phases austenitic or ferritic, a duplex mix of the two, the martensitic phase created when some steels are quenched from a high temperature, and precipitation-hardenable.

Ferritic Stainless Steel In ferritic stainless steel, the iron and chromium atoms are arranged in what is termed a body-centered cubic structure in which the atoms are arranged on the corners of the cube and one in the center (Figure E.1). As well as being ferro-magnetic, ferritic stainless steel exhibits very high stress corrosion-cracking resistance. Ferritic stainless steels are plain chromium (10.5 to 18%) grades ­characterized by moderate corrosion resistance and poor fabrication ­properties. These characteristics may be improved with the addition of molybdenum; some, aluminum or titanium. The basic 430 grade is a simple corrosion and heat-resisting grade. Alloying elements that tend to make the structure ferritic are called ­ferrite formers and result in grades such as Grades 434 and 444 and in the proprietary grade 3CR12. Common ferritic grades include: 18Cr-2Mo, 26Cr-1Mo, 29Cr-4Mo, and 29Cr-4Mo-2Ni (Figure E.2).

Austenitic Stainless Steel With the addition of nickel, the properties change dramatically. As shown (Figure E.3), the atoms are re-arranged so that they occur on the corners of the cube and also in the center of each of the faces. In this manner, it becomes what is termed austenitic stainless steel. Ferrite

Body centred cubic crystal Figure E.1.  In ferritic stainless steel, the atoms are arranged in a body-centered cubic structure.

Stainless Steel   •  257 444 Cr/Mo 17.7/2.1 Add more Mo for further improved corrosion resistance 434 Cr/Mo 17/1 Add Mo for improved corrosion resistance 430 Basic grade Cr 16.5

Added niobium for increased corrosion and heat resistance

436 Cr/Mo/Nb 17.5/1.2/0.6

Lower Cr plus Al prevents hardening when cooled from high temperatures. 405 Cr/Al 12/0.2 Lowest Cr 409 Cr/Ti 11/0.5

Figure E.2.  Ferritic stainless steel family.

Ferrite

Austenite

Add Nickel

Body centred cubic structure

Face centred cubic structure

Figure E.3.  With the addition of nickel, the atoms in austenitic stainless steels are arranged on the corners of the cube and also in the center of each of the faces.

It can, thus, be seen from Table E.1 that, unless you are specifically looking for a ferro-magnetic material, austenitic stainless steel would be the most obvious choice. Indeed, this is borne out by the fact that austenitic stainless steels account for about 70% or more of all stainless steel used worldwide, with ferritic stainless steels making up about 25%. The other families each represent less than 1% of the total market. Austenitic stainless steels are designated by numbers in the 200 and 300 series.

258  •  Stainless Steel

Table E.1.  Difference in the properties of ferritic and austenitic stainless steels. Properties Toughness Ductility Weldability Thermal expansion Stress corrosion cracking ­resistance Magnetic properties

Ferritic Moderate Moderate Limited Moderate Very high

Austenitic Very high Very high Good High Low

Ferro ­magnetic

Non-magnetic

Series 300 The relationship between the 300 austenitic grades is shown in Figure E.4 The basic grade 304 contains about 18% chromium and 8% nickel (often referred to as 18/8) and range through to the high alloy or ‘super austenitics’ such as 904L and 6% molybdenum grades. Ni-Cr-Fe Alloys Add Ni for corrosion resistance in high temperature applications

303, 303Se

Add S or Se for machinability

347

309, 310, 314, 330

Add Cr and Ni for strength and oxidation resistance

304 18Cr-8Ni

Add Nb + Ta to reduce sensitization

Add Ti to reduce sensitization

Add Mo for pitting resistance 316

304L Lower C to reduce sensitization

Add more Mo for more pitting resistance 317

321

316L 317L

Add Ni, Mo, N for corrosion resistance

Super austenitic stainless steel

Figure E.4.  The relationship between the various 300 series austenitic grades.

Stainless Steel   •  259

Additional elements can be added, such as molybdenum, titanium, or copper, to modify or improve their properties, making them suitable for many critical applications involving high temperature, as well as corrosion resistance. This group of steels is also suitable for cryogenic applications because the effect of the nickel content in making the steel austenitic avoids the problems of brittleness at low temperatures, which is a characteristic of other types of steel. Generally, the 300 grade alloys are subject to crevice and pitting corrosion. The time it takes for this type of corrosion to occur is called the ‘incubation time.’ In seawater, the incubation time for machinable grades, such as Type 303, is practically zero while that for the best alloys, such as Type 316, the time ranges from six months to a year. Low-carbon versions, (indicated by the letter suffix L) include 304L, 316L, and 317L, in which the carbon content of the alloy is below 0.03%. This reduces the effect of ‘sensitization’ in which chromium carbides precipitate at the grain boundaries due to the high temperatures involved in welding. The relatively high nickel content also inhibits the brittleness exhibited by ferritic materials at low temperatures, and thus makes austenitic steels suitable for cryogenic applications. 200 Series We have seen earlier how the addition of nickel is used in the creation of the classic chrome–nickel 300 series austenitic stainless steel. The reduced nickel content of the 200 series chrome–manganese grades makes them significantly cheaper. However, depending on their chemistry, they also offer good formability (ductility) and/or strength. Indeed, certain grades (201, 202, and 205 series) even offer about 30% higher yield strength than the classic 304-series chrome–nickel grade, allowing designers to cut weight (Table E.2). Reducing nickel, on the other hand, reduces the maximum chromium content possible in the alloy. Less chromium means less corrosion resistance and a consequent narrowing of the range of applications for which the material is suitable. A word of warning comes from the International Stainless Steel Forum (ISSF). Continuous pressure to cut costs, especially from the Asian market, has resulted in the development of austenitic grades ever lower in nickel and chromium, often not covered by international codes or specifications. In fact, numerous chrome–manganese grades are company-specific and identified simply by a title given to them by the producer.

260  •  Stainless Steel

Table E.2.  Chemical composition of standard grades (courtesy International Stainless Steel Forum) Chemical composition (wt. %) Mn Cr Ni N

Grades

C

Cu

201

0.15 max

5.50–7.50 16.0–18.0 3.50–5.50 0.25 max

-

202

0.15 max

7.50–10.0

17.0–19.0 4.00–6.00 0.25 max

-

204

0.15 max

6.50–9.0

15.5–17.5

205

0.12–0.25

14.0–15.5

16.5–18.0 1.0–1.75 0.32–0.40

1.5–3.5 0.05–0.25 2.0–4.0 -

Duplex Stainless Steels Duplex stainless steels are a mixture of austenite and ferrite microstructures that combine some of the features of each class: • • • •

resistance to stress corrosion cracking, but inferior to ferritic steel; superior toughness to ferritic steel, but inferior to austenitic steel; roughly twice the strength of austenitic steel; superior resistance to pitting, crevice corrosion, and stress corrosion cracking; • high resistance to chloride ions attack; and • high weldability. These features are achieved by adding less nickel than would be necessary for making a fully austenitic stainless steel. Thus, Grade 2304 comprises 23% chromium and 4% nickel, while Grade 2205 comprises 22% chromium and 5% nickel—with both grades containing further minor alloying additions. On the negative side, austenitic–ferritic duplex stainless steels are only usable between temperature limits of about –50°C and 300°C, outside which they suffer reduced toughness.

Martensitic Stainless Steel Named after the German metallurgist, Adolf Martens, the martensitic Grade 400 series (Figure E.5) are low-carbon (0.1–1%), low-nickel (less than 2%) steels containing chromium (12 to 14%) and molybdenum (0.2–1%).

Stainless Steel   •  261 414

416

Add Ni for improved corrosion resistance

410 Basic grade 0.15C-13.5Cr

Add P + S for improved machinability

Add C to improve mechanical properties

420 Increased C to improve toughness

Increased Cr for increased corrosion resistance

440

Figure E.5.  The martensitic Grade 400 series.

Stainless steels hardened by transformation to martensite are tempered to give the desired engineering properties. At high temperatures, they have an austenitic structure that is transformed into martensitic structure upon cooling to room temperature. Unfortunately, this tempering can influence corrosion susceptibility. For example, corrosion susceptibility of type 420 stainless steel is at its maximum when the alloy is ­tempered at temperatures in the range of 450° to 600°C. So, although not as ­corrosion-resistant as the 200 and 300 classes, martensitic stainless steels are magnetic, extremely strong (if not a little brittle), highly machinable, and can be hardened by heat treatment. Martensitic stainless steels are subject to both uniform and non-­ uniform attack in seawater. And, the incubation time for a non-uniform attack in even weak chlorides is often only a few hours or days. Precipitation-Hardening Martensitic Stainless Steels These chromium- and nickel-containing steels can be precipitation-­ hardened to develop very high tensile strengths. Precipitation-hardening stainless steels are usually designated by a trade name, rather than by their AISI 600 series designations. The most common grade in this group is ‘17-4 PH,’ also known as Grade 630, with a composition of 17% chromium, 4% nickel, 4% copper,

262  •  Stainless Steel

and 0.3% niobium. The main advantage of these steels is that they can be supplied in the ‘solution-treated’ condition, in which state the steel is just machinable. Following machining, forming, etc., the steel can be hardened by a single, fairly low-temperature ‘ageing’ heat treatment that causes no distortion of the component. Precipitation-hardening generally results in a slight increase in corrosion susceptibility and an increased susceptibility to hydrogen embrittlement. Figure E.6 shows the relationship between the complete family of stainless steels. Superferritic stainless steel

Ni-Cr-Fe alloys Add Ni for corrosion resistance in high temperature applications

Add Cr, Mo 430 347

No Ni, Ferritic Add Nb +Ta to reduce sensitization

321

316L 317L Superaustenitic stainless steel

Add S or Se for machinability

309, 310, 314, 330 Add Cr and Ni for Strength and oxidation resistance

Add Ti to reduce sensitization

304L

303, 303Se

304 Fe-19Cr-10Ni Add Mo for pitting resistance

Lower C to reduce sensitization Add Ni, Mo, N for corrosion resistance

316 Add more Mo for pitting resistance

317

Duplex stainless steel

Increase Cr, lower Ni for higher strength Precipitation hardening stainless steel

Add Cu, Ti, Al, lower Ni

Add Mn and N, lower Ni for higher strength No Ni addition, lower Cr, Martensitic

201, 202

403, 410, 420

Figure E.6.  Relationship between the complete family of stainless steels.

Glossary ∆P AChI ANSI API ASME ASTM AWG BSI CO CV DIN DN E/P FCI FL I/P IEC IEEE ISA ISO



J-T KV MAWP MOV MV NAMUR



Differential pressure American Chemical Institute American National Standards Institute American Petroleum Institute American Society of Mechanical Engineers American Society for Testing and Materials American Wire Gauge British Standards Institute Controller output Valve flow coefficient Deutsches Institit für Normung Nominal diameter Voltage to pneumatic converter Fluid Controls Institute Pressure recovery coefficient Current to pneumatic converter International Electrotechnical Commission Institute of Electrical and Electronic Engineers International Society for Automation International Organization for Standardization Note: ISO is not an acronym, but is based on the Greek word isos meaning equal. Joule–Thomson (effect) Valve flow coefficient (SI alternative = 0.865 × CV) Maximum allowable working pressure Motor-operated valve Manipulated variable Normen Arbeitsgen Mess Und Regeltechnik (loosely interpreted as Standards Work Group for Instruments and Controls.)

264  •  Glossary

NEMA OP PD PV PDR PN Q Qm Re SG SGf SGg SPL x XT Y Z



National Electrical Manufacturers Association Output Process demand Process variable Pressure drop ratio Nominal pressure Volumetric flow rate Mass flow rate Reynolds number Specific gravity Specific gravity of fluid Specific gravity of gas Sound pressure level Pressure drop ratio Choked value of pressure drop ratio Gas expansion factor Compressibility factor

Bibliography “Control Valve Trims and Devices to Control Cavitation Damage and Excessive Noise,” Mitech, Technical Product Bulletin No 1. “Introduction to Safety Valves.” Spirax Sarco, at: http://spiraxsarco.com/resources/ steam-engineering-tutorials/safety-valves/introduction-to-safety-valves.as “Pressure Relief Valve Engineering Handbook” Technical Document No. TP-V300, Crosby Valve Inc. “The Mitech Globe Control Valve Body,” Mitech, Technical Product Bulletin No 2. “Valve Signature Analysis” at: http://www2.emersonprocess.com/enUS/brands/ fisher/DigitalValveControllers/FIELDVUESolutions/ValveDiagnostics/ Pages/ValveSignatureBasics.aspx Bell, L.H., and D.H. Bell. 1994. Industrial Noise Control: Fundamentals and Applications. Marcel Dekker Inc. Boger, H., and L.Mazot. Why Most Control Valves Today are Throttling Around 60% Opening. Masoneilan-Dresser. Borden Jr., G. 1998. Control Valves. ISA. Campbell, J.M. February 2004. Gas Conditioning and Processing, Vol. 1: Basic Principles, 8th ed. Chris, W. 1999. A User’s Guide, Understanding Valve Actuators. Rotork Controls Inc. Chris, W. 2000. “New generation of valve actuators can provide important MOV Predictive Maintenance Data.” Rotork Controls Inc., Valve Magazine. Comparison of Different Valve Types. Crane Process Flow Technologies Ltd. Control Valve Noise Reduction. Fisher Rosemount. Dave, H. Understanding Control Valve Bench Set. Control Engineering. Dave, H. Understanding Control Valves. Control Engineering. Elonka, S., and A.R. Parsons. 1962. Standard Instrumentation Questions and Answers For Production-Processes Control, Vol. 1. McGraw-Hill. Emerson, G. 2005. Control Valve Handbook, 4th ed. Emerson Process Management. Herrmann, U.F. 1974. “Sound Reinforcement.” N.V. Philips’ Gloeilampenfabrieken, Eindhoven.

266  •  Bibliography Husu, M., I. Niemelä, J. Pyötsiä, and M. Simula. 1992. Flow Control Manual. Neles-Jamesbury. Hutchison, J.W. 1976. ISA Handbook of Control Valves. ISA. John, E. Positioner Guidelines. Emerson-Fisher-Rosemount. Mike Sessions, Cavitation Control in Control Valves. Practical Industrial Process Measurement for Engineers and Technicians. IDC Technologies. Richard, R. Designing a Positioner for the South African Market. Mitech. Sam, L. Control Valve Manual. Masoneilan. Stojkov, B.T. 1997. The Valve Primer. Industrial Press Inc.

About the Author Michael (Mick) Crabtree, Joining the Royal Air Force as an apprentice, Mick Crabtree trained in aircraft instrumentation and guided missiles. Completing his service career seconded to the Ministry of Defense as a technical writer, he emigrated to South Africa in 1966 where he worked, for many years, for a local manufacturing and systems integration company involved in industrial process control, SCADA, and PLC-based systems. Later, as an editor and managing editor of a leading monthly engineering journal, Mick wrote and published hundreds of articles, as well as eight technical resource handbooks on industrial process control: ‘Flow Measurement,’ ‘Temperature Measurement,’ ‘Analytical On-line Measurement,’ ‘Pressure and Level Measurement,’ ‘Valves,’ ‘Industrial Communications,’ and ‘The Complete Profibus Handbook.’ He subsequently founded his own PR and advertising company and was retained by a number of leading companies involved in the process control industry, including: Honeywell, Fisher-Rosemount, Krohne, Milltronics, and AEG. Apart from producing all their press releases and articles, he also undertook the conceptualization and production of a wide range of advertisements and data sheets, as well as newsletters. For the last 16 years, he has been involved in technical training and consultancy and has run workshops on industrial instrumentation and networking throughout the world (United States, Canada, United Kingdom, France, Southern Africa, Trinidad, Middle East, Australia, and New Zealand). During this period, he has led more than 6,000 engineers, ­technicians, and scientists on a variety of practical training workshops covering the fields of process control (loop tuning), process instrumentation, data communications, fieldbus, safety instrumentation systems (according to both ISA S84 and IEC 61508/61511), project management, online liquid analysis, and technical writing and communications.

268  •   About the Author

Completing his studies in Electrical, Electronic, and Instrumentation engineering, he holds an MSc in Industrial Flow Measurement from Huddersfield University. His hobbies and pastime include: cycling, rambling, history, and reading. After nearly 35 years spent in South Africa, he now lives in Wales, just outside Cardiff, having relocated to Britain some 18 years ago.

Index A Acoustics absolute levels, 236 decibel, 233–234 intensity, 231 logarithmic characteristic of ear, 231–233 pitch (frequency), 227–228 sound pressure level, 236–237 standards, 237–240 timbre, 228–229 velocity, 229–230 voltage ratios, 235–236 wavelength, 230 Actuators bench set, 183–184 cylinder actuator, 147–149 diaphragm actuator, 144–147 digital actuators, 155–156 electric actuation, 150–152 electro-hydraulic actuation, 149–150 flapper–nozzle assembly, 141–142 hammer-blow mechanism, 153, 154 I/P converter, 142–144 pneumatic control, 141 solenoid valve, 153–154 spool block, 149 torque limiting, 152–153 transfer mechanisms, 157–161

Alarm generation, digital positioners, 164 Austenitic stainless steel 300 series, 258–259 atom rearrangement, 256, 257 ferro-magnetic material, 257 B Balanced safety relief valves bellows-type, 204–206 limitations, 204 piston-type, 206–211 Bellows-type balanced safety relief valves back-pressure, 204 bellows failure, 205, 206 block schematics, 205 bonnet vents, 205–206 Bench Set, 183–184 Block-and-bleed system capital cost, 241 double, 241 function, 241 multi-block valve systems, 241 single double block-and-bleed manifold system, 241, 243 trunnion-mounted ball valves, 243 Buckling pin safety valve basic, 211, 212 buckling point, 213

270  •   Index

Euler’s law of compressed columns, 212–213 features, 213 stable conditions, 213 Burst disc advantages, 213–214 applications and installation practices, 218–219 configurations, 214–215 cyclic/pulsating duties, 221–223 holders, 214, 215 maximum operating pressure, 221, 222 performance tolerance, 219–221 C Composite rupture disks, 216–217 Conventional domed rupture disc, 215, 216 Conventional safety relief valves ASME/ANSI standard, 201 back-pressure API 520 Recommended Practice Guidelines, 203 built-up, 201 fluid inlet pressure, 202 forces acting on disc, 201 spring housing, 202, 203 superimposed, 200, 202–203 limitations, 200 spring housing, 201 Cyclic/pulsating duties, 221–223 Cylinder actuator cast cylinder, 147 vs. diaphragm cylinder, 147, 148 friction, 148–149 D Decibel, 233–234 Diaphragm actuator advantages, 147 air pressure failure, 144 direct-acting, 144, 145 disadvantages, 147

reverse-acting, 145, 146 springless, 145, 146 Digital actuators DC motor, 155 four-phase stepping motor, 156 stepping motor, 156 vernier arrangement, 156 Digital noise filtering, digital positioners, 164 Digital positioners, 164–165 Direct-acting diaphragm actuator, 144, 145 Direct-acting solenoid valve, 154, 155 Double-crank transfer mechanism advantage, 159 modulating control applications, 159 rocker plate movement, 158, 159 run torque, 158, 159 Duplex stainless steels, 260 E Electric actuators drawbacks, 151 non-reversing characteristics, 151 spring closure, 152 worm gear assembly, 151 Electro-hydraulic actuation drawback, 150 electronic control, 150 swing jet controller, 149, 150 Electronic torque monitoring, 176–177 F Ferritic stainless steel, 256 Flapper–nozzle assembly, 141–142 Fundamental frequency, 228 G Graphite rupture disc, 217–218

Index   •   271

H Hammer-blow mechanism, 153, 154 Harmonics, 228–229 High-pressure pilot-operated valve, 208 Hysteretic error, 168 I In-line repairs, 180, 182 Intensity, 231 J J–T valve construction, 225–226 Joule–Thomson effect, 225 reverse process, 225 K Kilogram, 184 L Low-pressure diaphragm-type pilot-operated valve, 209 M Maintenance-related data, digital positioner, 164 Martensitic stainless steel corrosion susceptibility, 261 grade 400 series, 260, 261 precipitation-hardening, 261–262 uniform and nonuniform attack, 261 N Newton, 184 Nitrogen-loaded surge relief valves, 252–253 Non-reclosing pressure safety relief valves buckling pin safety valve, 211–213

burst disc, 213–215 shear-pin safety valve, 211 P Pascal (Pa), 184 Passivation, 255 Pilot-operated safety relief valve advantages, 207 applications, 210 blockage, 210 high-pressure pilot-operated valve, 208 inlet pressure, 208 limitations, 210–211 low-pressure diaphragm-type, 209 overpressure and blowdown performance, 210 piston and seating arrangement, 208, 209 process pressure, 207 self-actuated auxiliary pressure relief valve, 207 Piston actuator. See Cylinder actuator Piston-type balanced safety valves back-pressure, 207 force balancing, 206, 207 pilot-operated safety relief valve, 207–210 Pitch (frequency), 227–228 Positioners digital, 164–165 electronic positioners, 162–163 feedback positioner, 162 I/P transducers, 161–162 principle of operation, 162 proportional control system, 161 set-point characterization, 163 Precipitation-hardening martensitic stainless steel, 261–262 Pressure, 184 Pressure safety relief valves applications, 194

272  •   Index

history, 187–190 limitations, 194 non-reclosing, 211–215 spring-loaded relief valves, 192–194 weight-loaded valves, 191–192 Pulsation dampener, 251–252 Pulsations, 250 R Rack and pinion transfer mechanism disadvantages, 157 double-piston arrangement, 158 Relief valves, 190 Reverse-acting diaphragm actuator, 145, 146 Reverse Joule–Thomson effect, 225 S Safety relief valves, 190. See also Pressure safety relief valves balanced, 204 composite rupture disks, 216–217 conventional, 200–203 conventional domed rupture disc, 215, 216 graphite rupture disc, 217–218 scored tension-loaded rupture discs, 217, 218 standards, 223–224 Safety valves closing pressure, 195 curtain area, 196 discharge area, 196 flow area, 195, 196 lifting, 196–198 reseating, 198–200 static inlet pressure, 194–195 Scored tension-loaded rupture discs, 217, 218

Scotch yoke transfer mechanism, 160–161 Shear-pin safety valve, 211 Shop repairs, 182–183 Signature analysis minimum and maximum friction value, 175 opening and closing lines, 174 packing friction, 175 revealed faults, 176 valve packing, 174 ‘valve signature’ plot, 174 Solenoid valve direct-acting, 154, 155 shut-off applications, 154 three-way solenoid valve, 155 Sound pressure level (SPL), 236–237 Spring calculations, 184–186 Springless diaphragm actuator, 145, 146 Springless piston-type actuator, 149 Spring-loaded pressure relief valves closed bonnets, 193, 194 closing force, 193 elements, 192–193 static inlet pressure, 193 static pressure, 192 Stainless steel 200 series, 259–260 300 grade alloys, 258–259 austenitic stainless steel, 256–260 classification issues, 255–256 duplex, 260 ferritic stainless steel, 256 martensitic, 260–262 passivation, 255 Stick-slip response, 170 Superimposed back-pressure, 200 Swing jet controller, 149, 150

Index   •   273

T Threshold of hearing, 237, 238 Threshold of pain, 237, 238 Timbre, 228–229 Torque limiting, 152–153 Transfer mechanisms double-crank mechanism, 158–160 rack and pinion, 157–158 scotch yoke mechanism, 160–161 Trunnion-mounted ball valves, 243 U Unbalance of force, 186 V Valve maintenance and repair actuator bench set, 183–184 drained system repair, 181 dynamic environment, 179 fluid leakage, 179 in-line repairs, 180, 182 packing replacement, 181 repairs under pressure, 180–181 seat rings replacement, 181–182 shaft leakage, 179 shop repairs, 182–183 spring calculations, 184–186 Valve stroke speed control, digital positioner, 165 Valve testing and diagnostics complete assembly, 171–173 deadband and hysteresis acceptable limits, 170 friction, 168–169 gear-train system, 167, 168

hysteretic error, 168 linear system, 167 step changes, 169 electronic torque monitoring, 176–177 non-linearity, 170–171 online diagnostics modern Fieldbus communication systems, 173 signature analysis, 174–176 stick-slip response, 170 Velocity, 229–230 Voltage ratios, 235–236 W Water hammer bulk modulus, 245 effective bulk modulus, 246–247 magnitude of pressure spike, 248 nitrogen-loaded surge relief valves, 252–253 prevention and mitigation, 250–251 pulsation dampener, 251–252 pulsations, 250 quick-closing valve, 245 sonic velocities, 245–246 steam condensate, 249–250 valve size and system condition, 248 velocity of pressure wave, 245, 246 Wavelength, 230 Weight-loaded pressure/vacuum relief valves, 191–192

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The Concise Valve Handbook Actuation, Maintenance, and Safety Relief, Volume II Michael A. Crabtree

CRABTREE

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AUTOMATION AND CONTROL COLLECTION

Research studies within the process industry routinely indicate that the fluid control valve is responsible for 60 to 70% of poorfunctioning control systems. Furthermore, valves in general are consistently wrongly selected, regularly misapplied, and often incorrectly installed. This two-volume book comprises a comprehensive up-to-date body of knowledge that provides a total in-depth insight into valve

THE CONTENT

and actuator technology—looking not just at control valves, but a

• Manufacturing Engineering • Mechanical & Chemical Engineering • Materials Science & Engineering • Civil & Environmental Engineering • Advanced Energy Technologies

whole host of other types including: check valves, shut-off valves, solenoid valves, and pressure relief valves.

Whilst studying the correct procedures for sizing, readers will also learn the correct procedures for calculating the spring ‘wind-up’ or ‘bench set’. Maintenance issues also include: testing for deadband/ hysteresis, stick-slip and non-linearity; on-line diagnostics; and signature analysis. Written in a detailed but understandable language, the two volumes are presented in a form suitable for both the beginner, with no prior knowledge of the subject, and the more advanced

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specialist. For the last sixteen years, ‘Mick’ Crabtree, who holds an MSc in industrial flow measurement, has been involved in technical training and consultancy—running workshops on industrial instrumentation and networking throughout the world covering the fields of process control (loop tuning), process instrumentation, data communications,

The Concise Valve Handbook, Volume II

A methodology is presented to ensure the optimum selection of size, choice of body and trim materials, components, and ancillaries.

The Concise Valve Handbook Actuation, Maintenance, and Safety Relief Volume II

fieldbus, safety instrumentation systems (according to both ISA S84 and IEC 61508/61511), project management, on-line analysis, and technical writing and communications. This book represents some thirty years wealth of experiential knowledge gleaned by the author working within a wide variety of industries and from more than 6000 technicians and engineers who have attended the author’s workshops. ISBN: 978-1-94708-369-1

Michael A. Crabtree