Page 1
Te c h n i c a l
The Technical Reference section includes articles covering regulator theory,
sizing, selection, overpressure protection, and other topics relating to
regulators. This section begins with the basic theory of regulators and ends
with conversion tables and other informative charts.
This section is for general reference only. For more detailed information
please visit www.emersonprocess.com/regulators or contact your
local Sales Ofce.
Page 2
Table of Contents
572
Te c h n i c a l
Regulator Control Theory
Fundamentals of Gas Pressure Regulators ................................. 577
Pilot-Operated Regulators .......................................................... 578
Conclusion .................................................................................. 578
Regulator Components
Straight Stem Style Direct-Operated .......................................... 579
Lever Style Direct-Operated ....................................................... 580
Loading Style (Two-Path Control) Pilot-Operated ..................... 581
Unloading Style Pilot-Operated .................................................... 582
Introduction to Regulators
Specic Regulator Types ............................................................ 583
Pressure Reducing Regulators ............................................ 583
Backpressure Regulators and Relief Valves ....................... 584
Pressure Switching Valves .................................................. 584
Vaccum Regulators and Breakers .......................................... 584
Types of Regulators .................................................................... 583
Direct-Operated (Self-Operated) Regulators ...................... 583
Pilot-Operated Regulators .................................................. 583
Regulator Selection Criteria ....................................................... 584
Control Application ............................................................ 585
Pressure Reducing Regulator Selection .............................. 585
Outlet Pressure to be Maintained ....................................... 585
Inlet Pressure of the Regulator ........................................... 585
Capacity Required .............................................................. 585
Shutoff Capability .............................................................. 585
Process Fluid ...................................................................... 585
Process Fluid Temperature ................................................. 585
Accuracy Required ............................................................ 586
Pipe Size Required ......................................................... 586
End Connection Style .................................................... 586
Required Materials ......................................................... 586
Control Lines .................................................................. 586
Stroking Speeds .............................................................. 586
Overpressure Protection ................................................ 586
Regulator Replacement ................................................. 586
Regulator Price ................................................................ 587
Backpressure Regulator Selection ................................... 587
Relief Valve Selection ............................................................ 587
Theory
Principles of Direct-Operated Regulators
Introduction ................................................................................ 588
Regulator Basics ......................................................................... 588
Essential Elements ...................................................................... 588
Restricting Element ............................................................ 589
Measuring Element ............................................................ 589
Loading Element ................................................................ 589
Regulator Operation ................................................................... 589
Increasing Demand ............................................................ 589
Decreasing Demand ........................................................... 589
Weights versus Springs ...................................................... 589
Spring Rate ......................................................................... 590
Equilibrium with a Spring .................................................. 590
Spring as Loading Element ........................................................ 590
Throttling Example ............................................................ 590
Regulator Operation and P2 ............................................... 591
Regulator Performance ....................................................... 591
Performance Criteria .......................................................... 591
Setpoint ............................................................................... 591
Droop .................................................................................. 591
Capacity .............................................................................. 591
Accuracy ............................................................................. 591
Lockup ................................................................................ 591
Spring Rate and Regulator Accuracy ......................................... 592
Spring Rate and Droop ....................................................... 592
Effect on Plug Travel .......................................................... 592
Light Spring Rate ............................................................... 592
Practical Limits................................................................... 592
Diaphragm Area and Regulator Accuracy ................................. 592
Diaphragm Area ................................................................. 592
Increasing Diaphragm Area ................................................ 593
Diaphragm Size and Sensitivity ......................................... 593
Restricting Element and Regulator Performance ....................... 593
Critical Flow ....................................................................... 593
Orice Size and Capacity ................................................... 594
Orice Size and Stability .................................................... 594
Orice Size, Lockup, and Wear .......................................... 594
Orice Guideline ................................................................ 594
Increasing P1 ....................................................................... 594
Factors Affecting Regulator Accuracy ....................................... 594
Performance Limits .................................................................... 594
Cycling ............................................................................... 594
Design Variations ................................................................ 594
Improving Regulator Accuracy with a Pitot Tube .............. 595
Numerical Example ............................................................ 595
Decreased Droop (Boost) ................................................... 595
Improving Performance with a Lever ........................................ 595
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Table of Contents
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Te c h n i c a l
Overpressure Protection Methods
Methods of Overpressure Protection .................................. 604
Relief Valves ........................................................................... 604
Types of Relief Valves ................................................... 604
Advantages ...................................................................... 604
Disadvantages ................................................................ 604
Monitoring Regulators .......................................................... 605
Advantages ...................................................................... 605
Disadvantages ................................................................. 605
Working Monitor .................................................................... 605
Series Regulation .................................................................... 605
Advantages ...................................................................... 606
Disadvantages ................................................................. 606
Shutoff Devices ...................................................................... 606
Advantages ...................................................................... 606
Disadvantages ........................................................... 606
Relief Monitor ................................................................. 606
Summary .......................................................................... 607
Principles of Relief Valves
Overpressure Protection .................................................. 608
Maximum Pressure Considerations ................................. 608
Downstream Equipment ........................................... 608
Main Regulator ......................................................... 608
Piping ....................................................................... 608
Relief Valves ................................................................... 608
Relief Valve Popularity ............................................ 609
Relief Valve Types .................................................... 609
Selection Criteria ............................................................. 609
Pressure Build-up ...................................................... 609
Periodic Maintenance ............................................... 609
Cost versus Performance .......................................... 609
Installation and Maintenance Considerations .......... 609
Pop Type Relief Valve ..................................................... 609
Operation .................................................................. 609
Typical Applications ................................................. 610
Advantages ............................................................... 610
Disadvantage ............................................................ 610
Direct-Operated Relief Valves .................................. 610
Operation .................................................................. 610
Product Example ...................................................... 611
Typical Applications ................................................. 611
Selection Criteria ...................................................... 611
Pilot-Operated Relief Valves ........................................... 612
Operation .................................................................. 612
Product Example ...................................................... 612
Performance ............................................................. 613
Typical Applications ................................................. 613
Selection Criteria ...................................................... 613
Principles of Pilot-Operated Regulators
Pilot-Operated Regulator Basics ................................................ 596
Regulator Pilots .................................................................. 596
Gain .................................................................................... 596
Identifying Pilots ................................................................ 596
Setpoint ............................................................................... 596
Spring Action ...................................................................... 596
Pilot Advantage .................................................................. 596
Gain and Restrictions ................................................................. 596
Stability .............................................................................. 596
Restrictions, Response Time, and Gain ............................. 597
Loading and Unloading Designs ................................................ 597
Two-Path Control (Loading Design) .................................. 597
Two-Path Control Advantages ............................................ 598
Unloading Control .............................................................. 598
Unloading Control Advantages .......................................... 598
Performance Summary ............................................................... 598
Accuracy ............................................................................. 598
Capacity .............................................................................. 598
Lockup ................................................................................ 599
Applications ....................................................................... 599
Two-Path Control ....................................................................... 599
Type 1098-EGR .................................................................. 599
Type 99 ............................................................................... 600
Unloading Design ....................................................................... 600
Selecting and Sizing Pressure Reducing Regulators
Introduction ................................................................................ 601
Quick Selection Guides ................................................. 601
Product Pages .................................................................. 601
The Role of Experience ................................................. 601
Special Requirements .................................................... 601
Sizing Equations ..................................................................... 601
General Sizing Guidelines .................................................... 602
Body Size ......................................................................... 602
Construction .................................................................... 602
Pressure Ratings ............................................................. 602
Wide-Open Flow Rate ................................................... 602
Outlet Pressure Ranges and Springs............................ 602
Accuracy .......................................................................... 602
Inlet Pressure Losses ...................................................... 602
Orice Diameter ............................................................. 602
Speed of Response ......................................................... 602
Turn-Down Ratio ............................................................ 602
Sizing Exercise: Industrial Plant Gas Supply ................... 602
Quick Selection Guide ................................................ 603
Product Pages ................................................. 603
Final Selection ................................................................ 603
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Table of Contents
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Te c h n i c a l
Vacuum Control
Vacuum Applications ................................................................. 620
Vacuum Terminology ......................................................... 620
Vacuum Control Devices ................................................... 620
Vacuum Regulators ................................................................... 620
Vacuum Breakers (Relief Valves) .............................................. 620
Vacuum Regulator Installation Examples .................................. 622
Vacuum Breaker Installation Examples .................................. 623
Gas Blanketing in Vacuum ......................................................... 625
Features of Fisher® Brand Vacuum Regulators and Breakers ... 625
Valve Sizing Calculations (Traditional Method)
Introduction ..................................................................... 626
Sizing for Liquid Service ................................................ 626
Viscosity Corrections ............................................... 626
Finding Valve Size ................................................... 626
Nomograph Instructions ................................................... 627
Nomograph Equations ................................................... 627
Nomograph Procedure ................................................... 627
Predicting Flow Rate ............................................... 628
Predicting Pressure Drop .......................................... 628
Flashing and Cavitation ............................................ 628
Choked Flow ............................................................ 629
Liquid Sizing Summary ........................................... 631
Liquid Sizing Nomenclature .................................... 631
Sizing for Gas or Steam Service ...................................... 632
Universal Gas Sizing Equation ................................ 632
General Adaptation for Steam and Vapors ............... 633
Special Equation for Steam Below 1000 psig .......... 633
Gas and Steam Sizing Summary .............................. 634
Valve Sizing (Standardized Method)
Introduction ....................................................... 635
Liquid Sizing ............................................................... 635
Sizing Valves for Liquids ............................................... 635
Liquid Sizing Sample Problem ...................................... 638
Gas and Steam Sizing ................................................... 641
Sizing Valves for Compressible Fluids ......................... 641
Compressible Fluid Sizing Sample Problem ................ 642
Temperature Considerations
Cold Temperature Considerations
Regulators Rated for Low Temperatures ......................... 647
Selection Criteria ............................................................. 647
Internal Relief .................................................................. 613
Operation .................................................................. 613
Product Example ...................................................... 613
Performance and Typical Applications ..................... 614
Selection Criteria ...................................................... 614
Selection and Sizing Criteria ........................................... 614
Maximum Allowable Pressure ................................. 614
Regulator Ratings ..................................................... 614
Piping ....................................................................... 614
Maximum Allowable System Pressure .................... 614
Determining Required Relief Valve Flow ................ 614
Determine Constant Demand ................................... 615
Selecting Relief Valves .................................................... 615
Required Information ............................................... 615
Regulator Lockup Pressure ...................................... 615
Identify Appropriate Relief Valves ........................... 615
Final Selection .......................................................... 615
Applicable Regulations ............................................ 615
Sizing and Selection Exercise ......................................... 615
Initial Parameters ..................................................... 615
Performance Considerations .................................... 615
Upstream Regulator .................................................. 615
Pressure Limits ......................................................... 615
Relief Valve Flow Capacity ...................................... 615
Relief Valve Selection .............................................. 616
Principles of Series Regulation and Monitor Regulators
Series Regulation ............................................................. 617
Failed System Response ........................................... 617
Regulator Considerations ......................................... 617
Applications and Limitations ................................... 617
Upstream Wide-Open Monitors ...................................... 617
System Values ........................................................... 617
Normal Operation ..................................................... 617
Worker Regulator B Fails ......................................... 618
Equipment Considerations ....................................... 618
Downstream Wide-Open Monitors .................................. 618
Normal Operation ..................................................... 618
Worker Regulator A Fails ......................................... 618
Upstream Versus Downstream Monitors ......................... 618
Working Monitors ............................................................ 618
Downstream Regulator ..................................................... 619
Upstream Regulator ......................................... 619
Normal Operation ..................................................... 619
Downstream Regulator Fails ......................................... 619
Upstream Regulator Fails ..................................................... 619
Sizing Monitor Regulators .............................................. 619
Estimating Flow when Pressure Drop is Critical ..... 619
Assuming P
intermediate
to Determine Flow .................... 619
Fisher Monitor Sizing Program ....................................... 619
Page 5
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Te c h n i c a l
Freezing
Introduction ..................................................................... 648
Reducing Freezing Problems ........................................... 648
Heat the Gas ............................................................. 648
Antifreeze Solution .................................................. 648
Equipment Selection ................................................ 648
System Design .......................................................... 649
Water Removal ......................................................... 649
Summary .......................................................................... 649
Sulfide Stress Cracking—NACE MR0175-2002,
MR0175/ISO 15156
The Details ....................................................................... 650
New Sulfide Stress Cracking Standards For Refineries .... 651
Responsibility .................................................................. 651
Applicability of NACE MR0175/ISO 15156 .................. 651
Basics of Sulde Stress Cracking (SSC) and
Stress Corrosion Cracking (SCC) ......................................... 651
Carbon Steel .................................................................... 652
Carbon and Low-Alloy Steel
Welding Hardness Requirements ....................................... 653
Low-Alloy Steel Welding Hardness Requirements .. 653
Cast Iron .......................................................................... 653
Stainless Steel .................................................................. 653
400 Series Stainless Steel ......................................... 653
300 Series Stainless Steel ......................................... 653
S20910 ...................................................................... 654
CK3MCuN ............................................................... 654
S17400 ...................................................................... 654
Duplex Stainless Steel .............................................. 654
Highly Alloyed Austenitic Stainless Steels .............. 654
Nonferrous Alloys ........................................................... 655
Nickel-Base Alloys .................................................. 655
Monel® K500 and Inconel® X750 ........................ 655
Cobalt-Base Alloys ................................................... 655
Aluminum and Copper Alloys .................................. 656
Titanium ................................................................... 656
Zirconium ................................................................. 656
Springs ............................................................................ 656
Coatings ........................................................................... 656
Stress Relieving ............................................................... 656
Bolting ............................................................................. 656
Bolting Coatings .............................................................. 657
Composition Materials .................................................... 657
Tubulars ............................................................................ 657
Expanded Limits and Materials ....................................... 657
Codes and Standards ........................................................ 658
Certifications ................................................................... 658
Reference
Chemical Compatibility of Elastomers and Metals
Introduction ..................................................................... 659
Elastomers: Chemical Names and Uses ......................... 659
General Properties of Elastomers .................................... 660
Fluid Compatibility of Elastomers .................................. 661
Compatibility of Metals ................................................... 662
Regulator Tips
Regulator Tips ................................................................. 664
Conversions, Equivalents, and Physical Data
Pressure Equivalents ....................................................... 666
Pressure Conversion - Pounds per Square Inch
(PSI) to Bar ....................................................... 666
Volume Equivalents ......................................................... 666
Volume Rate Equivalents ................................................ 667
Mass Conversion—Pounds to Kilograms ....................... 667
Temperature Conversion Formulas ................................. 667
Area Equivalents ............................................................. 667
Kinematic-Viscosity Conversion Formulas ..................... 667
Conversion Units ............................................................. 668
Other Useful Conversions ............................................... 668
Converting Volumes of Gas ............................................ 668
Fractional Inches to Millimeters ...................................... 669
Length Equivalents .......................................................... 669
Whole Inch-Millimeter Equivalents ................................ 669
Metric Prefixes and Symbols ........................................... 669
Greek Alphabet ................................................................ 669
Length Equivalents - Fractional and Decimal
Inches to Millimeters ........................................ 670
Temperature Conversions ................................................ 671
A.P.I. and Baumé Gravity Tables and Weight Factors ..... 674
Characteristics of the Elements ....................................... 675
Recommended Standard Specications for Valve
Materials Pressure-Containing Castings ........... 676
Physical Constants of Hydrocarbons ............................... 679
Physical Constants of Various Fluids .............................. 680
Properties of Water .......................................................... 682
Properties of Saturated Steam ......................................... 682
Properties of Saturated Steam—Metric Units ................. 685
Properties of Superheated Steam ..................................... 686
Determine Velocity of Steam in Pipes ............................. 689
Recommended Steam Pipe Line Velocities ..................... 689
Typical Condensation Rates in Insulated Pipes ............... 689
Typical Condensation Rates without Insulation .............. 689
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576
Te c h n i c a l
Flow of Water Through Schedule 40 Steel Pipes ............ 690
Flow of Air Through Schedule 40 Steel Pipes ................ 692
Average Properties of Propane ........................................ 694
Orifice Capacities for Propane ......................................... 694
Standard Domestic Propane Tank Specifications ............ 694
Approximate Vaporization Capacities
of Propane Tanks ............................................... 694
Pipe and Tubing Sizing .................................................... 695
Vapor Pressures of Propane ............................................. 695
Converting Volumes of Gas ............................................. 695
BTU Comparisons ......................................................... 695
Capacities of Spuds and Orifices ..................................... 696
Kinematic Viscosity - Centistokes ................................... 699
Specic Gravity of Typical Fluids vs
Temperature ................................................................ 700
Effect of Inlet Swage On Critical Flow
Cg Requirements ......................................................... 701
Seat Leakage Classications ...................................................... 702
Nominal Port Diameter and Leak Rate ...................................... 702
Flange, Valve Size, and Pressure-Temperature Rating
Designations ............................................................... 703
Equivalency Table .......................................................... 704
Pressure-Temperature Ratings for Valve Bodies ............. 704
ASME Face-To-Face Dimensions for
Flanged Regulators ..................................................... 706
Diameter of Bolt Circles ................................................. 706
Wear and Galling Resistance Chart of Material
Combinations ............................................................. 707
Equivalent Lengths of Pipe Fittings and Valves ......................... 707
Pipe Data: Carbon and Alloy Steel —Stainless Steel ..... 708
American Pipe Flange Dimensions ................................ 710
EN 1092-1 Cast Steel Flange Standards .................................. 710
EN 1092-1 Pressure/Temperature Ratings for
Cast Steel Valve Ratings ............................................... 711
Drill Sizes for Pipe Taps ................................................. 712
Standard Twist Drill Sizes .............................................. 712
Glossary
Glossary of Terms ...................................................................... 713
Page 7
Regulator Control Theory
577
Te c h n i c a l
Flow
P
2
P
1
Restricting Element
Load
Flow
Regulator
Flow
Load
Regulator
P
2
Diaphragm
Spring
= Loading Pressure
Flow
P
1
P
2
P
L
Diaphragm
Spring
= Loading Pressure
Flow
P
L
P
2
P
1
Fundamentals of Gas Pressure Regulators
The primary function of any gas regulator is to match the ow of
gas through the regulator to the demand for gas placed upon the
system. At the same time, the regulator must maintain the system
pressure within certain acceptable limits.
A typical gas pressure system might be similar to that shown in
Figure 1, where the regulator is placed upstream of the valve or
other device that is varying its demand for gas from the regulator.
The loading element can be one of any number of things such as
a weight, a hand jack, a spring, a diaphragm actuator, or a piston
actuator, to name a few of the more common ones.
A diaphragm actuator and a spring are frequently combined, as
shown in Figure 3, to form the most common type of loading
element. A loading pressure is applied to a diaphragm to produce
a loading force that will act to close the restricting element. The
spring provides a reverse loading force which acts to overcome
the weight of the moving parts and to provide a fail-safe operating
action that is more positive than a pressure force.
If the load ow decreases, the regulator ow must decrease also.
Otherwise, the regulator would put too much gas into the system,
and the pressure (P2) would tend to increase. On the other hand, if
the load ow increases, then the regulator ow must increase also
in order to keep P2 from decreasing due to a shortage of gas in the
pressure system.
From this simple system it is easy to see that the prime job of the
regulator is to put exactly as much gas into the piping system as the
load device takes out.
If the regulator were capable of instantaneously matching its
ow to the load ow, then we would never have major transient
variation in the pressure (P2) as the load changes rapidly. From
practical experience we all know that this is normally not the
case, and in most real-life applications, we would expect some
uctuations in P2 whenever the load changes abruptly.
Because the regulator’s job is to modulate the ow of gas into
the system, we can see that one of the essential elements of any
regulator is a restricting element that will t into the ow stream
and provide a variable restriction that can modulate the ow of gas
through the regulator.
Figure 2 shows a schematic of a typical regulator restricting
element. This restricting element is usually some type of valve
arrangement. It can be a single-port globe valve, a cage style
valve, buttery valve, or any other type of valve that is capable of
operating as a variable restriction to the ow.
In order to cause this restricting element to vary, some type of
loading force will have to be applied to it. Thus we see that the
second essential element of a gas regulator is a Loading Element
that can apply the needed force to the restricting element.
Figure 1
So far, we have a restricting element to modulate the ow through
the regulator, and we have a loading element that can apply the
necessary force to operate the restricting element. But, how do
we know when we are modulating the gas ow correctly? How
do we know when we have the regulator ow matched to the load
ow? It is rather obvious that we need some type of Measuring
Element which will tell us when these two ows have been
perfectly matched. If we had some economical method of directly
measuring these ows, we could use that approach; however, this
is not a very feasible method.
We noted earlier in our discussion of Figure 1 that the system
pressure (P2) was directly related to the matching of the two ows.
If the restricting element allows too much gas into the system, P2
will increase. If the restricting element allows too little gas into
the system, P2 will decrease. We can use this convenient fact to
provide a simple means of measuring whether or not the regulator
is providing the proper ow.
Figure 2
Figure 3
Page 8
Regulator Control Theory
578
Te c h n i c a l
Flow
P
1
P
2
P
2
P
1
Manometers, Bourdon tubes, bellows, pressure gauges, and
diaphragms are some of the possible measuring elements that
we might use. Depending upon what we wish to accomplish,
some of these measuring elements would be more advantageous
than others. The diaphragm, for instance, will not only act as a
measuring element which responds to changes in the measured
pressure, but it also acts simultaneously as a loading element. As
such, it produces a force to operate the restricting element that
varies in response to changes in the measured pressure. If we
add this typical measuring element to the loading element and the
restricting element that we selected earlier, we will have a complete
gas pressure regulator as shown in Figure 4.
Let’s review the action of this regulator. If the restricting element
tries to put too much gas into the system, the pressure (P2) will
increase. The diaphragm, as a measuring element, responds to this
increase in pressure and, as a loading element, produces a force
which compresses the spring and thereby restricts the amount
of gas going into the system. On the other hand, if the regulator
doesn’t put enough gas into the system, the pressure (P2) falls and
the diaphragm responds by producing less force. The spring will
then overcome the reduced diaphragm force and open the valve
to allow more gas into the system. This type of self-correcting
action is known as negative feedback. This example illustrates
that there are three essential elements needed to make any
operating gas pressure regulator. They are a restricting element,
a loading element, and a measuring element. Regardless of how
sophisticated the system may become, it still must contain these
three essential elements.
Pilot-Operated Regulators
So far we have only discussed direct-operated regulators.
This is the name given to that class of regulators where the
measured pressure is applied directly to the loading element
with no intermediate hardware. There are really only two basic
congurations of direct-operated regulators that are practical.
These two basic types are illustrated in Figures 4 and 5.
Figure 4
Figure 5
If the proportional band of a given direct-operated regulator is
too great for a particular application, there are a number of things
we can do. From our previous examples we recall that spring
rate, valve travel, and effective diaphragm area were the three
parameters that affect the proportional band. In the last section we
pointed out the way to change these parameters in order to improve
the proportional band. If these changes are either inadequate or
impractical, the next logical step is to install a pressure amplier in
the measuring or sensing line. This pressure amplier is frequently
referred to as a pilot.
Conclusion
It should be obvious at this point that there are fundamentals to
understand in order to properly select and apply a gas regulator
to do a specic job. Although these fundamentals are profuse
in number and have a sound theoretical base, they are relatively
straightforward and easy to understand.
As you are probably aware by now, we made a number of
simplifying assumptions as we progressed. This was done in the
interest of gaining a clearer understanding of these fundamentals
without getting bogged down in special details and exceptions. By
no means has the complete story of gas pressure regulation been
told. The subject of gas pressure regulation is much broader in
scope than can be presented in a single document such as this, but
it is sincerely hoped that this application guide will help to gain
a working knowledge of some fundamentals that will enable one
to do a better job of designing, selecting, applying, evaluating, or
troubleshooting any gas pressure regulation equipment.
Page 9
Regulator Components
579
Te c h n i c a l
Straight Stem Style Direct-Operated Regulator Components
Type 133L
ORIFICE
INLET PRESSURE
BOOST PRESSURE
OUTLET PRESSURE
ATMOSPHERIC PRESSURE
SPRING SEAT
CAGE
VALVE STEM
BALANCING
DIAPHRAGM
SPRING ADJUSTOR
SPRING CASE
DIAPHRAGM
HEAD
VALVE DISK
CLOSING CAP
CONTROL SPRING
VENT
EXTERNAL
CONTROL LINE
CONNECTION
REGISTRATION
DIAPHRAGM
BODY
DIAPHRAGM
CASE
NOTE:
THE INFORMATION PRESENTED IS FOR REFERENCE ONLY. FOR MORE SPECIFIC APPLICATION
INFORMATION, PLEASE LOG ON TO: www.emersonprocess.com/regulators
A6555
Page 10
Regulator Components
580
Te c h n i c a l
Lever Style Direct-Operated Regulator Components
Type 627
BODY
VALVE DISK
LEVER
VENT
VALVE STEM
SPRING SEAT
SPRING
DIAPHRAGM
ORIFICE
DIAPHRAGM
HEAD
DIAPHRAGM
CASE
CLOSING CAP
SPRING CASE
INTERNAL CONTROL
REGISTRATION
ADJUSTING SCREW
INLET PRESSURE
OUTLET PRESSURE
ATMOSPHERIC PRESSURE
NOTE:
THE INFORMATION PRESENTED IS FOR REFERENCE ONLY. FOR MORE SPECIFIC APPLICATION
INFORMATION, PLEASE LOG ON TO: www.emersonprocess.com/regulators
A6557
Page 11
Regulator Components
581
Te c h n i c a l
Loading Style (Two-Path Control) Pilot-Operated Regulator Components
Type 1098-EGR
A6563
Type 1098-EGR
TRAVEL INDICATOR
CONTROL LINE
(EXTERNAL)
TYPE 1098 ACTUATOR
PILOT
INLET PRESSURE
INLINE FILTER
VENT
PILOT SUPPLY
PRESSURE
MAIN SPRING
BODY
BALANCED EGR TRIM
INLET PRESSURE
OUTLET PRESSURE
LOADING PRESSURE
ATMOSPHERIC PRESSURE
NOTE:
THE INFORMATION PRESENTED IS FOR REFERENCE ONLY. FOR MORE SPECIFIC APPLICATION
INFORMATION, PLEASE LOG ON TO: www.emersonprocess.com/regulators
A6563
Page 12
Regulator Components
582
Te c h n i c a l
Unloading Style Pilot-Operated Regulator Components
Type EZR
TYPE 161EB PILOT
TYPE 252 PILOT
SUPPLY FILTER
TYPE 112 RESTRICTOR
STRAINER
BODY
VENT
EXTERNAL
CONTROL
LINE
TRAVEL
INDICATOR
PLUG AND
DIAPHRAGM
TRIM
CAGE
EXTERNAL PILOT
SUPPLY LINE
INLET PRESSURE
OUTLET PRESSURE
LOADING PRESSURE
ATMOSPHERIC PRESSURE
NOTE:
THE INFORMATION PRESENTED IS FOR REFERENCE ONLY. FOR MORE SPECIFIC APPLICATION
INFORMATION, PLEASE LOG ON TO: www.emersonprocess.com/regulators
W7438
Page 13
Introduction to Regulators
583
Te c h n i c a l
Instrument engineers agree that the simpler a system is the
better it is, as long as it provides adequate control. In general,
regulators are simpler devices than control valves. Regulators are
self-contained, direct-operated control devices which use energy
from the controlled system to operate whereas control valves
require external power sources, transmitting instruments, and
control instruments.
Specific Regulator Types
Within the broad categories of direct-operated and pilotoperated regulators fall virtually all of the general regulator
designs, including:
• Pressure reducing regulators
• Backpressure regulators
• Pressure relief valves
• Pressure switching valves
• Vacuum regulators and breakers
Pressure Reducing Regulators
A pressure reducing regulator maintains a desired reduced outlet
pressure while providing the required uid ow to satisfy a
downstream demand. The pressure which the regulator maintains
is the outlet pressure setting (setpoint) of the regulator.
Types of Pressure Reducing Regulators
This section describes the various types of regulators. All
regulators t into one of the following two categories:
1. Direct-Operated (also sometimes called Self-Operated)
2. Pilot-Operated
Direct-Operated (Self-Operated) Regulators
Direct-operated regulators are the simplest style of regulators. At
low set pressures, typically below 1 psig (0,07 bar), they can have
very accurate (±1%) control. At high control pressures, up to
500 psig (34,5 bar), 10 to 20% control is typical.
In operation, a direct-operated, pressure reducing regulator
senses the downstream pressure through either internal pressure
registration or an external control line. This downstream pressure
opposes a spring which moves the diaphragm and valve plug to
change the size of the ow path through the regulator.
Pilot-Operated Regulators
Pilot-operated regulators are preferred for high ow rates or where
precise pressure control is required. A popular type of pilotoperated system uses two-path control. In two-path control, the
main valve diaphragm responds quickly to downstream pressure
Figure 1. Type 627 Direct-Operated Regulator and Operational Schematic
Figure 2. Type 1098-EGR Pilot-Operated Regulator and Operational Schematic
INLET PRESSURE
OUTLET PRESSURE
ATMOSPHERIC PRESSURE
INLET PRESSURE
OUTLET PRESSURE
LOADING PRESSURE
ATMOSPHERIC PRESSURE
Type 1098-EGR
W6956
A6563
A6557
W4793
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Introduction to Regulators
584
Te c h n i c a l
changes, causing an immediate correction in the main valve plug
position. At the same time, the pilot diaphragm diverts some
of the reduced inlet pressure to the other side of the main valve
diaphragm to control the nal positioning of the main valve plug.
Two-path control results in fast response and accurate control.
Backpressure Regulators and Pressure Relief Valves
A backpressure regulator maintains a desired upstream pressure
by varying the ow in response to changes in upstream pressure.
A pressure relief valve limits pressure build-up (prevents
overpressure) at its location in a pressure system. The relief valve
opens to prevent a rise of internal pressure in excess of a specied
value. The pressure at which the relief valve begins to open
pressure is the relief pressure setting.
Relief valves and backpressure regulators are the same devices.
The name is determined by the application. Fisher® relief valves
are not ASME safety relief valves.
Vacuum Regulators and Breakers
Vacuum regulators and vacuum breakers are devices used to
control vacuum. A vacuum regulator maintains a constant vacuum
at the regulator inlet with a higher vacuum connected to the outlet.
During operation, a vacuum regulator remains closed until a
vacuum decrease (a rise in absolute pressure) exceeds the spring
setting and opens the valve disk. A vacuum breaker prevents a
vacuum from exceeding a specied value. During operation, a
vacuum breaker remains closed until an increase in vacuum
(a decrease in absolute pressure) exceeds the spring setting and
opens the valve disk.
Regulator Selection Criteria
This section describes the procedure normally used to select
regulators for various applications. For most applications, there
is generally a wide choice of regulators that will accomplish the
Figure 3. Type 63EG Backpressure Regulator/Relief Valve
Operational Schematic
Figure 4. Type Y690VB Vacuum Breaker and Type V695VR Vacuum Regulator
Operational Schematics
Pressure Switching Valves
Pressure switching valves are used in pneumatic logic systems. These
valves are for either two-way or three-way switching. Two-way
switching valves are used for on/off service in pneumatic systems.
Three-way switching valves direct inlet pressure from one outlet
port to another whenever the sensed pressure exceeds or drops
below a preset limit.
INLET PRESSURE
OUTLET PRESSURE
ATMOSPHERIC PRESSURE
LOADING PRESSURE
INLET PRESSURE
CONTROL PRESSURE (VACUUM)
ATMOSPHERIC PRESSURE
VACUUM
BEING CONTROLLED
HIGHER
VACUUM SOURCE
VACUUM
PUMP
VACUUM
PUMP
VACUUM
BEING LIMITED
A6929
B2583
B2582
TYPE Y690VB
TYPE Y695VR
Page 15
Introduction to Regulators
585
Te c h n i c a l
required function. The vendor and the customer, working together,
have the task of deciding which of the available regulators is best
suited for the job at hand. The selection procedure is essentially a
process of elimination wherein the answers to a series of questions
narrow the choice down to a specic regulator.
Control Application
To begin the selection procedure, it’s necessary to dene what
the regulator is going to do. In other words, what is the control
application? The answer to this question will determine the general
type of regulator required, such as:
• Pressure reducing regulators
• Backpressure regulators
• Pressure relief valves
• Vacuum regulators
• Vacuum breaker
The selection criteria used in selecting each of these general regulator
types is described in greater detail in the following subsections.
Pressure Reducing Regulator Selection
The majority of applications require a pressure reducing regulator.
Assuming the application calls for a pressure reducing regulator,
the following parameters must be determined:
• Outlet pressure to be controlled
• Inlet pressure to the regulator
• Capacity required
• Shutoff capability required
• Process uid
• Process uid temperature
• Accuracy required
• Pipe size required
• End connection style
• Material requirements
• Control line needed
• Overpressure protection
Outlet Pressure to be Controlled
For a pressure reducing regulator, the rst parameter to determine
is the required outlet pressure. When the outlet pressure is known,
it helps determine:
• Spring requirements
• Casing pressure rating
• Body outlet rating
• Orice rating and size
• Regulator size
Inlet Pressure of the Regulator
The next parameter is the inlet pressure. The inlet pressure
(minimum and maximum) determines the:
• Pressure rating for the body inlet
• Orice pressure rating and size
• Main spring (in a pilot-operated regulator)
• Regulator size
If the inlet pressure varies signicantly, it can have an effect on:
• Accuracy of the controlled pressure
• Capacity of the regulator
• Regulator style (two-stage or unloading)
Capacity Required
The required ow capacity inuences the following decisions:
• Size of the regulator
• Orice size
• Style of regulator (direct-operated or pilot-operated)
Shutoff Capability
The required shutoff capability determines the type of
disk material:
• Standard disk materials are Nitrile (NBR) and Neoprene (CR),
these materials provide the tightest shutoff.
• Other materials, such as Nylon (PA), Polytetrauoroethylene
(PTFE), Fluoroelastomer (FKM), and Ethylenepropylene
(EPDM), are used when standard material cannot be used.
• Metal disks are used in high temperatures and when
elastomers are not compatible with the process uid;
however, tight shutoff is typically not achieved.
Process Fluid
Each process uid has its own set of unique characteristics in
terms of its chemical composition, corrosive properties, impurities,
ammability, hazardous nature, toxic effect, explosive limits, and
molecular structure. In some cases special care must be taken
to select the proper materials that will come in contact with the
process uid.
Process Fluid Temperature
Fluid temperature might determine the materials used in the
regulator. Standard regulators use Steel and Nitrile (NBR) or
Neoprene (CR) elastomers that are good for a temperature range
of -40° to 180°F (-40° to 82°C). Temperatures above and below
this range may require other materials, such as Stainless steel,
Ethylenepropylene (EPDM), or Peruoroelastomer (FFKM).
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Introduction to Regulators
586
Te c h n i c a l
Accuracy Required
The accuracy requirement of the process determines the acceptable
droop (also called proportional band or offset). Regulators fall into
the following groups as far as droop is concerned:
• Rough-cut Group— This group generally includes many
rst-stage, rough-cut direct-operated regulators. This group
usually has the highest amount of droop. However, some
designs are very accurate, especially the low-pressure gas or
air types, such as house service regulators, which incorporate
a relatively large diaphragm casing.
• Close-control Group— This group usually includes pilot-
operated regulators. They provide high accuracy over a
large range of ows. Applications that require close control
include these examples:
• Burner control where the fuel/air ratio is critical to
burner efciency and the gas pressure has a signicant
effect on the fuel/air ratio.
• Metering devices, such as gas meters, which require
constant input pressures to ensure accurate measurement.
Pipe Size Required
If the pipe size is known, it gives the specier of a new regulator
a more dened starting point. If, after making an initial selection
of a regulator, the regulator is larger than the pipe size, it usually
means that an error has been made either in selecting the pipe size
or the regulator, or in determining the original parameters (such
as pressure or ow) required for regulator selection. In many
cases, the outlet piping needs to be larger than the regulator for the
regulator to reach full capacity.
End Connection Style
In general, the following end connections are available for the
indicated regulator sizes:
• Pipe threads or socket weld: 2-inch (DN 50) and smaller
• Flanged: 1-inch (DN 25) and larger
• Butt weld: 1-inch (DN 25) and larger
Note: Not all end connections are available for all regulators.
Required Materials
The regulator construction materials are generally dictated by the
application. Standard materials are:
• Aluminum
• Cast iron or Ductile iron
• Steel
• Bronze and Brass
• Stainless steel
Special materials required by the process can have an effect on the
type of regulator that can be used. Oxygen service, for example,
requires special materials, requires special cleaning preparation,
and requires that no oil or grease be in the regulator.
Control Lines
For pressure registration, control lines are connected downstream
of a pressure reducing regulator, and upstream of a backpressure
regulator. Typically large direct-operated regulators have
external control lines, and small direct-operated regulators have
internal registration instead of a control line. Most pilot-operated
regulators have external control lines, but this should be conrmed
for each regulator type considered.
Stroking Speed
Stroking speed is often an important selection criteria. Directoperated regulators are very fast, and pilot-operated regulators
are slightly slower. Both types are faster than most control valves.
When speed is critical, techniques can be used to decrease
stroking time.
Overpressure Protection
The need for overpressure protection should always be considered.
Overpressure protection is generally provided by an external relief
valve, or in some regulators, by an internal relief valve. Internal
relief is an option that you must choose at the time of purchase.
The capacity of internal relief is usually limited in comparison with
a separate relief valve. Other methods such as shutoff valves or
monitor regulators can also be used.
Regulator Replacement
When a regulator is being selected to replace an existing regulator,
the existing regulator can provide the following information:
• Style of regulator
• Size of regulator
• Type number of the regulator
• Special requirements for the regulator, such as downstream
pressure sensing through a control line versus internal
pressure registration.
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Introduction to Regulators
587
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Figure 5. Backpressure Regulator/Relief Valve Applications
RELIEF PRESSURE CONTROL AT RELIEF VALVE INLET BACKPRESSURE CONTROL
Regulator Price
The price of a regulator is only a part of the cost of ownership.
Additional costs include installation and maintenance. In selecting
a regulator, you should consider all of the costs that will accrue over
the life of the regulator. The regulator with a low initial cost might
not be the most economical in the long run. For example, a directoperated regulator is generally less expensive, but a pilot-operated
regulator might provide more capacity for the initial investment.
To illustrate, a 2-inch (DN 50) pilot-operated regulator can have
the same capacity and a lower price than a 3-inch (DN 80), direct-
operated regulator.
Backpressure Regulator Selection
Backpressure regulators control the inlet pressure rather than the
outlet pressure. The selection criteria for a backpressure regulator
the same as for a pressure reducing regulator.
Relief Valve Selection
An external relief valve is a form of backpressure regulator. A
relief valve opens when the inlet pressure exceeds a set value.
Relief is generally to atmosphere. The selection criteria is the same
as for a pressure reducing regulator.
MAIN VALVE
VENT VALVE B
BLOCK VALVE A
MAIN PRESSURE LINE
NONRESTRICTIVE
VENTS AND PIPING
ALTERNATE PILOT
EXHAUST PIPING
PILOT
CONTROL LINE
30B8289_A
BLOCK VALVE A
VENT VALVE B
MAIN VALVE
VENT VALVE D
ALTERNATE PILOT
EXHAUST PIPING
VENT VALVE C
PILOT
CONTROL LINE
30B8288_A
Page 18
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588
Te c h n i c a l
Introduction
Pressure regulators have become very familiar items over the years,
and nearly everyone has grown accustomed to seeing them in
factories, public buildings, by the roadside, and even on the outside
of their own homes. As is frequently the case with such familiar
items, we have a tendency to take them for granted. It’s only
when a problem develops, or when we are selecting a regulator
for a new application, that we need to look more deeply into the
fundamentals of the regulator’s operation.
Regulators provide a means of controlling the ow of a gas or
other uid supply to downstream processes or customers. An
ideal regulator would supply downstream demand while keeping
downstream pressure constant; however, the mechanics of direct-
operated regulator construction are such that there will always be
some deviation (droop or offset) in downstream pressure.
The service regulator mounted on the meter outside virtually every
home serves as an example. As appliances such as a furnace or
stove call for the ow of more gas, the service regulator responds
by delivering the required ow. As this happens, the pressure
should be held constant. This is important because the gas meter,
which is the cash register of the system, is often calibrated for a
given pressure.
Direct-operated regulators have many commercial and residential
uses. Typical applications include industrial, commercial, and
domestic gas service, instrument air supply, and a broad range of
applications in industrial processes.
Regulators automatically adjust ow to meet downstream demand.
Before regulators were invented, someone had to watch a pressure
gauge for pressure drops which signaled an increase in downstream
demand. When the downstream pressure decreased, more ow was
required. The operator then opened the regulating valve until the
gauge pressure increased, showing that downstream demand was
being met.
Essential Elements
Direct-operated regulators have three essential elements:
• A restricting element — a valve, disk, or plug
• A measuring element— generally a diaphragm
• A loading element— generally a spring
Figure 1. Direct-Operated Regulators
TYPE 630
Regulator Basics
A pressure reducing regulator must satisfy a downstream demand
while maintaining the system pressure within certain acceptable
limits. When the ow rate is low, the regulator plug or disk
approaches its seat and restricts the ow. When demand increases,
the plug or disk moves away from its seat, creating a larger
opening and increased ow. Ideally, a regulator should provide a
constant downstream pressure while delivering the required ow.
Figure 2. Three Essential Elements
TYPE HSR
133 SERIES
MEASURING
ELEMENT
LOADING
ELEMENT
(WEIGHT)
RESTRICTING
ELEMENT
W1327
W1934
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Principles of Direct-Operated Regulators
589
Te c h n i c a l
Restricting Element
The regulator’s restricting element is generally a disk or plug that
can be positioned fully open, fully closed, or somewhere in between
to control the amount of ow. When fully closed, the disk or plug
seats tightly against the valve orice or seat ring to shutoff ow.
Measuring Element
The measuring element is usually a exible diaphragm that
senses downstream pressure (P2). The diaphragm moves
as pressure beneath it changes. The restricting element is
often attached to the diaphragm with a stem so that when the
diaphragm moves, so does the restricting element.
Loading Element
A weight or spring acts as the loading element. The loading
element counterbalances downstream pressure (P2). The amount of
unbalance between the loading element and the measuring element
determines the position of the restricting element. Therefore, we can
adjust the desired amount of ow through the regulator, or setpoint,
by varying the load. Some of the rst direct-operated regulators used
weights as loading elements. Most modern regulators use springs.
Regulator Operation
To examine how the regulator works, let’s consider these values for
a direct-operated regulator installation:
• Upstream Pressure (P
1
) = 100 psig
• Downstream Pressure (P
2
) = 10 psig
• Pressure Drop Across the Regulator (P) = 90 psi
• Diaphragm Area (A
D
) = 10 square inches
• Loading Weight = 100 pounds
Let’s examine a regulator in equilibrium as shown in Figure 3. The
pressure acting against the diaphragm creates a force acting up to
100 pounds.
Diaphragm Force (FD) = Pressure (P2) x Area of Diaphragm (AD)
or
FD = 10 psig x 10 square inches = 100 pounds
The 100 pounds weight acts down with a force of 100 pounds,
so all the opposing forces are equal, and the regulator plug
remains stationary.
Increasing Demand
If the downstream demand increases, P2 will drop. The pressure on
the diaphragm drops, allowing the regulator to open further. Suppose
in our example P2 drops to 9 psig. The force acting up then equals
90 pounds (9 psig x 10 square inches = 90 pounds). Because of the
unbalance of the measuring element and the loading element, the
restricting element will move to allow passage of more ow.
Decreasing Demand
If the downstream demand for ow decreases, downstream
pressure increases. In our example, suppose P2 increases to 11
psig. The force acting up against the weight becomes 110 pounds
(11 psig x 10 square inches = 110 pounds). In this case, unbalance
causes the restricting element to move up to pass less ow or
lockup.
Weights versus Springs
One of the problems with weight-loaded systems is that they are
slow to respond. So if downstream pressure changes rapidly,
our weight-loaded regulator may not be able to keep up. Always
behind, it may become unstable and cycle—continuously going
from the fully open to the fully closed position. There are other
problems. Because the amount of weight controls regulator
setpoint, the regulator is not easy to adjust. The weight will always
have to be on top of the diaphragm. So, let’s consider using a
spring. By using a spring instead of a weight, regulator stability
increases because a spring has less stiffness.
AT EQUILIBRIUM
Figure 3. Elements
100 LB
AREA = 10 IN
2
P2 = 10 PSIGP1 = 100 PSIG
FW = 100 LB
FD = 100 LB
FD = (P2 x AD) = (10 PSIG)(10 IN2) = 100 LB
OPEN
100 LB
AREA = 10 IN
2
P2 = 9 PSIGP1 = 100 PSIG
FW = 100 LB
FD = 90 LB
FD = (P2 x AD) = (9 PSIG)(10 IN2) = 90 LB
Page 20
Principles of Direct-Operated Regulators
590
Te c h n i c a l
Spring Rate
We choose a spring for a regulator by its spring rate (K). K
represents the amount of force necessary to compress the spring
one inch. For example, a spring with a rate of 100 pounds per inch
needs 100 pounds of force to compress it one inch, 200 pounds of
force to compress it two inches, and so on.
Equilibrium with a Spring
Instead of a weight, let’s substitute a spring with a rate of
100 pounds per inch. And, with the regulator’s spring adjustor, we’ll
wind in one inch of compression to provide a spring force (FS) of
100 pounds. This amount of compression of the regulator spring
determines setpoint, or the downstream pressure that we want to
hold constant. By adjusting the initial spring compression, we
change the spring loading force, so P2 will be at a different value in
order to balance the spring force.
Now the spring acts down with a force of 100 pounds, and the
downstream pressure acts up against the diaphragm producing
a force of 100 pounds (FD = P2 x AD). Under these conditions
the regulator has achieved equilibrium; that is, the plug or disk is
holding a xed position.
Spring as Loading Element
By using a spring instead of a xed weight, we gain better control
and stability in the regulator. The regulator will now be less likely to
go fully open or fully closed for any change in downstream pressure
(P2). In effect, the spring acts like a multitude of different weights.
Throttling Example
Assume we still want to maintain 10 psig downstream. Consider
what happens now when downstream demand increases and
pressure P2 drops to 9 psig. The diaphragm force (FD) acting up is
now 90 pounds.
FD = P2 x AD
FD = 9 psig x 10 in
2
FD = 90 pounds
We can also determine how much the spring will move (extend)
which will also tell us how much the disk will travel. To keep the
regulator in equilibrium, the spring must produce a force (FS) equal
to the force of the diaphragm. The formula for determining spring
force (FS) is:
FS = (K)(X)
where K = spring rate in pounds/inch and X = travel or
compression in inches.
AREA = 10 IN
2
SPRING AS ELEMENT
FS = F
D
FS = (K)(X)
FD = (P2)(AD)
F
S
F
D
(TO KEEP DIAPHRAGM
FROM MOVING)
FD = (10 PSIG)(102) = 100 LB
FS = (100 LB/IN)(X) = 100 LB
X = 1-INCH COMPRESSION
FS = 90 LB
FD = 90 LB
P1 = 100 PSIG P2 = 9 PSIG
Figure 4. Spring as Element
Figure 5. Plug Travel
0.1-INCH
AT EQUILIBRIUM
P1 = 100 PSIG P2 = 10 PSIG
Page 21
Principles of Direct-Operated Regulators
591
Te c h n i c a l
We know FS is 90 pounds and K is 100 pounds/inch, so we can
solve for X with:
X = FS ÷ K
X = 90 pounds ÷ 100 pounds/inch
X = 0.9 inch
The spring, and therefore the disk, has moved down 1/10-inch,
allowing more ow to pass through the regulator body.
Regulator Operation and P
2
Now we see the irony in this regulator design. We recall that the
purpose of an ideal regulator is to match downstream demand
while keeping P2 constant. But for this regulator design to increase
ow, there must be a change in P2.
Regulator Performance
We can check the performance of any regulating system by
examining its characteristics. Most of these characteristics can
be best described using pressure versus ow curves as shown in
Figure 6.
Performance Criteria
We can plot the performance of an ideal regulator such that
no matter how the demand changes, our regulator will match
that demand (within its capacity limits) with no change in the
downstream pressure (P2). This straight line performance becomes
the standard against which we can measure the performance of a
real regulator.
Setpoint
The constant pressure desired is represented by the setpoint. But
no regulator is ideal. The downward sloping line on the diagram
represents pressure (P2) plotted as a function of ow for an actual
direct-operated regulator. The setpoint is determined by the initial
compression of the regulator spring. By adjusting the initial spring
compression you change the spring loading force, so P2 will be
at a different value in order to balance the spring force. This
establishes setpoint.
Droop
Droop, proportional band, and offset are terms used to describe
the phenomenon of P2 dropping below setpoint as ow increases.
Droop is the amount of deviation from setpoint at a given ow,
expressed as a percentage of setpoint. This “droop” curve is
important to a user because it indicates regulating (useful) capacity.
Capacity
Capacities published by regulator manufacturers are given for
different amounts of droop. Let’s see why this is important.
Let’s say that for our original problem, with the regulator set at
10 psig, our process requires 200 SCFH (standard cubic feet per
hour) with no more than a 1 psi drop in setpoint. We need to keep
the pressure at or above 9 psig because we have a low limit safety
switch set at 9 psig that will shut the system down if pressure falls
below this point.
Figure 6 illustrates the performance of a regulator that can do the job.
And, if we can allow the downstream pressure to drop below 9 psig,
the regulator can allow even more ow.
The capacities of a regulator published by manufacturers are
generally given for 10% droop and 20% droop. In our example,
this would relate to ow at 9 psig and at 8 psig.
Accuracy
The accuracy of a regulator is determined by the amount of ow it
can pass for a given amount of droop. The closer the regulator is to
the ideal regulator curve (setpoint), the more accurate it is.
Lockup
Lockup is the pressure above setpoint that is required to shut the
regulator off tight. In many regulators, the orice has a knife
edge while the disk is a soft material. Some extra pressure, P2, is
Figure 6. Typical Performance Curve
AS THE FLOW RATE APPROACHES ZERO, P2 INCREASES STEEPLY. LOCKUP IS THE TERM APPLIED TO THE
VALUE OF P2 AT ZERO FLOW.
LOCKUP
SETPOINT
DROOP
(OFFSET)
WIDE-OPEN
P
2
FLOW
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Principles of Direct-Operated Regulators
592
Te c h n i c a l
required to force the soft disk into the knife edge to make a tight
seal. The amount of extra pressure required is lockup pressure.
Lockup pressure may be important for a number of reasons.
Consider the example above where a low pressure limit switch
would shut down the system if P2 fell below 9 psig. Now consider
the same system with a high pressure safety cut out switch set a
10.5 psig. Because our regulator has a lockup pressure of 11 psig,
the high limit switch will shut the system down before the regulator
can establish tight shutoff. Obviously, we’ll want to select a
regulator with a lower lockup pressure.
Spring Rate and Regulator Accuracy
Using our initial problem as an example, let’s say we now need the
regulator to ow 300 SCFH at a droop of 10% from our original
setpoint of 10 psig. Ten percent of 10 psig = 1 psig, so P2 cannot
drop below 10 to 1, or 9 psi. Our present regulator would not be
accurate enough. For our regulator to pass 300 SCFH, P2 will have
to drop to 8 psig, or 20% droop.
Spring Rate and Droop
One way to make our regulator more accurate is to change to
a lighter spring rate. To see how spring rate affects regulator
accuracy, let’s return to our original example. We rst tried a
spring with a rate of 100 pounds/inch. Let’s substitute one with a
rate of 50 pounds/inch. To keep the regulator in equilibrium, we’ll
have to initially adjust the spring to balance the 100 pound force
produced by P2 acting on the diaphragm. Recall how we calculate
spring force:
FS = K (spring rate) x X (compression)
Knowing that FS must equal 100 pounds and K = 50 pounds/inch,
we can solve for X, or spring compression, with:
X = FS ÷ K, or X = 2 inches
So, we must wind in 2-inches of initial spring compression to
balance diaphragm force, FD.
Effect on Plug Travel
We saw before that with a spring rate of 100 pounds/inch, when
P2 dropped from 10 to 9 psig, the spring relaxed (and the valve
disk traveled) 0.1 inch. Now let’s solve for the amount of disk
travel with the lighter spring rate of 50 pounds per inch. The force
produced by the diaphragm is still 90 pounds.
FD = P2 x AD
To maintain equilibrium, the spring must also produce a force of
90 pounds. Recall the formula that determines spring force:
FS = (K)(X)
Because we know FS must equal 90 pounds and our spring rate (K)
is 50 pounds/inch, we can solve for compression (X) with:
X = FS ÷ K
X = 90 pounds ÷ 50 pounds/inch
X = 1.8 inches
To establish setpoint, we originally compressed this spring 2 inches.
Now it has relaxed so that it is only compressed 1.8 inches, a change
of 0.2-inch. So with a spring rate of 50 pounds/inch, the regulator
responded to a 1 psig drop in P2 by opening twice as far as it did
with a spring rate of 100 pounds/inch. Therefore, our regulator
is now more accurate because it has greater capacity for the same
change in P2. In other words, it has less droop or offset. Using this
example, it is easy to see how capacity and accuracy are related and
how they are related to spring rate.
Light Spring Rate
Experience has shown that choosing the lightest available spring
rate will provide the most accuracy (least droop). For example, a
spring with a range of 35 to 100 psig is more accurate than a spring
with a range of 90 to 200 psig. If you want to set your regulator at
100 psig, the 35 to 100 psig spring will provide better accuracy.
Practical Limits
While a lighter spring can reduce droop and improve accuracy,
using too light a spring can cause instability problems. Fortunately,
most of the work in spring selection is done by regulator
manufacturers. They determine spring rates that will provide good
performance for a given regulator, and publish these rates along
with other sizing information.
Diaphragm Area and Regulator Accuracy
Diaphragm Area
Until this point, we have assumed the diaphragm area to be
constant. In practice, the diaphragm area changes with travel.
We’re interested in this changing area because it has a major
inuence on accuracy and droop.
Diaphragms have convolutions in them so that they are exible
enough to move over a rated travel range. As they change position,
Page 23
Principles of Direct-Operated Regulators
593
Te c h n i c a l
they also change shape because of the pressure applied to them.
Consider the example shown in Figure 7. As downstream pressure
(P2) drops, the diaphragm moves down. As it moves down, it
changes shape and diaphragm area increases because the centers of
the convolutions become further apart. The larger diaphragm area
magnies the effect of P2 so even less P2 is required to hold the
diaphragm in place. This is called diaphragm effect. The result is
decreased accuracy because incremental changes in P2 do not result
in corresponding changes in spring compression or disk position.
Increasing Diaphragm Area
To better understand the effects of changing diaphragm area, let’s
calculate the forces in the exaggerated example given in Figure 7.
First, assume that the regulator is in equilibrium with a downstream
pressure P2 of 10 psig. Also assume that the area of the diaphragm
in this position is 10 square inches. The diaphragm force (FD) is:
FD = (P2)(AD)
FD = (10 psi) (10 square inches)
FD = 100 pounds
Now assume that downstream pressure drops to 9 psig signaling the
need for increased ow. As the diaphragm moves, its area increases
to 11 square inches. The diaphragm force now produced is:
FD = (9 psi) (11 square inches)
FD = 99 pounds
The change in diaphragm area increases the regulator’s droop.
While it’s important to note that diaphragm effect contributes to
Figure 8. Critical Flow
droop, diaphragm sizes are generally determined by manufacturers
for different regulator types, so there is rarely a user option.
Diaphragm Size and Sensitivity
Also of interest is the fact that increasing diaphragm size can result
in increased sensitivity. A larger diaphragm area will produce
more force for a given change in P2. Therefore, larger diaphragms
are often used when measuring small changes in low-pressure
applications. Service regulators used in domestic gas service are
an example.
Restricting Element and Regulator Performance
Critical Flow
Although changing the orice size can increase capacity, a
regulator can pass only so much ow for a given orice size and
inlet pressure, no matter how much we improve the unit’s accuracy.
Shown in Figure 8, after the regulator is wide-open, reducing P2
does not result in higher ow. This area of the ow curve identies
critical ow. To increase the amount of ow through the regulator,
the owing uid must pass at higher and higher velocities. But, the
uid can only go so fast. Holding P1 constant while decreasing P2,
ow approaches a maximum which is the speed of sound in that
particular gas, or its sonic velocity. Sonic velocity depends on the
inlet pressure and temperature for the owing uid. Critical ow is
generally anticipated when downstream pressure (P2) approaches a
value that is less than or equal to one-half of inlet pressure (P1).
Figure 7. Changing Diaphragm Area
DIAPHRAGM EFFECT ON DROOP
A = 10 IN
2
F
D1
A = 11 IN
2
F
D2
WHEN P2 DROPS TO 9
FD = P2 x A
FD1 = 10 x 10 = 100 LB
FD2 = 9 x 11 = 99 LB
SETPOINT
WIDE-OPEN
FLOW
P
2
CRITICAL FLOW
Page 24
Principles of Direct-Operated Regulators
594
Te c h n i c a l
Orifice Size and Capacity
One way to increase capacity is to increase the size of the orice.
The variable ow area between disk and orice depends directly
on orice diameter. Therefore, the disk will not have to travel
as far with a larger orice to establish the required regulator
ow rate, and droop is reduced. Sonic velocity is still a limiting
factor, but the ow rate at sonic velocity is greater because more
gas is passing through the larger orice.
Stated another way, a given change in P2 will produce a larger
change in ow rate with a larger orice than it would with a smaller
orice. However, there are denite limits to the size of orice that
can be used. Too large an orice makes the regulator more sensitive
to uctuating inlet pressures. If the regulator is overly sensitive, it
will have a tendency to become unstable and cycle.
Orifice Size and Stability
One condition that results from an oversized orice is known as
the “bathtub stopper” effect. As the disk gets very close to the
orice, the forces of uid ow tend to slam the disk into the orice
and shutoff ow. Downstream pressure drops and the disk opens.
This causes the regulator to cycle—open, closed, open, closed. By
selecting a smaller orice, the disk will operate farther away from
the orice so the regulator will be more stable.
Orifice Size, Lockup, and Wear
A larger orice size also requires a higher shutoff pressure,
or lockup pressure. In addition, an oversized orice usually
produces faster wear on the valve disk and orice because it
controls ow with the disk near the seat. This wear is accelerated
with high ow rates and when there is dirt or other erosive
material in the ow stream.
Orifice Guideline
Experience indicates that using the smallest possible orice is
generally the best rule-of-thumb for proper control and stability.
Increasing P1
Regulator capacity can be increased by increasing inlet pressure (P1).
Factors Affecting Regulator Accuracy
As we have seen, the design elements of a regulator—the spring,
diaphragm, and orice size—can affect its accuracy. Some
of these inherent limits can be overcome with changes to the
regulator design.
Performance Limits
The three curves in Figure 9 summarize the effects of spring rate,
diaphragm area, and orice size on the shape of the controlled
pressure-ow rate curve. Curve A is a reference curve representing
Figure 9. Increased Sensitivity
a typical regulator. Curve B represents the improved performance
from either increasing diaphragm area or decreasing spring rate.
Curve C represents the effect of increasing orice size. Note that
increased orice size also offers higher ow capabilities. But
remember that too large an orice size can produce problems that
will negate any gains in capacity.
Cycling
The sine wave in Figure 10 might be what we see if we increase
regulator sensitivity beyond certain limits. The sine wave indicates
instability and cycling.
Design Variations
All direct-operated regulators have performance limits that result
from droop. Some regulators are available with features designed
to overcome or minimize these limits.
FLOW
SETPOINT
A
B
C
P
2
Figure 10. Cycling
SETPOINT
TIME
P
2
FLOW INCREASED
Page 25
Principles of Direct-Operated Regulators
595
Te c h n i c a l
E0908
Type HSR
Improving Regulator Accuracy with a Pitot Tube
In addition to the changes we can make to diaphragm area, spring
rate, orice size, and inlet pressure, we can also improve regulator
accuracy by adding a pitot tube as shown in Figure 11. Internal to
the regulator, the pitot tube connects the diaphragm casing with a
low-pressure, high velocity region within the regulator body. The
pressure at this area will be lower than P2 further downstream.
By using a pitot tube to measure the lower pressure, the regulator
will make more dramatic changes in response to any change in
P2. In other words, the pitot tube tricks the regulator, causing it to
respond more than it would otherwise.
Improving Performance with a Lever
The lever style regulator is a variation of the simple direct-operated
regulator. It operates in the same manner, except that it uses a lever
to gain mechanical advantage and provide a high shutoff force.
In earlier discussions, we noted that the use of a larger diaphragm
can result in increased sensitivity. This is because any change in P2
will result in a larger change in diaphragm force. The same result
is obtained by using a lever to multiply the force produced by the
diaphragm as shown in Figure 13.
The main advantage of lever designs is that they provide increased
force for lockup without the extra cost, size, and weight associated
with larger diaphragms, diaphragm casings, and associated parts.
Figure 11. Pitot Tube
P2 = 10 PSIGP1 = 100 PSIG
SENSE PRESSURE HERE
Decreased Droop (Boost)
The pitot tube offers one chief advantage for regulator accuracy, it
decreases droop. Shown in Figure 12, the diaphragm pressure, PD,
must drop just as low with a pitot tube as without to move the disk
far enough to supply the required ow. But the solid curve shows
that P2 does not decrease as much as it did without a pitot tube. In
fact, P2 may increase. This is called boost instead of droop. So
the use of a pitot tube, or similar device, can dramatically improve
droop characteristics of a regulator.
Figure 13. Lever Style Regulator
Figure 12. Performance with Pitot Tube
PRESSURE
SETPOINT
P
2
P
D
FLOW
P2 = DOWNSTREAM PRESSURE
PD = PRESSURE UNDER DIAPHRAGM
Numerical Example
For example, we’ll establish setpoint by placing a gauge
downstream and adjusting spring compression until the gauge
reads 10 psig for P2. Because of the pitot tube, the regulator might
actually be sensing a lower pressure. When P2 drops from 10 psig
to 9 psig, the pressure sensed by the pitot tube may drop from
8 psig to 6 psig. Therefore, the regulator opens further than it
would if it were sensing actual downstream pressure.
PIVOT POINT
INLET PRESSURE
OUTLET PRESSURE
ATMOSPHERIC PRESSURE
E0908
Page 26
Principles of Pilot-Operated Regulators
596
Te c h n i c a l
Pilot-Operated Regulator Basics
In the evolution of pressure regulator designs, the shortcomings of
the direct-operated regulator naturally led to attempts to improve
accuracy and capacity. A logical next step in regulator design is to
use what we know about regulator operation to explore a method
of increasing sensitivity that will improve all of the performance
criteria discussed.
Identifying Pilots
Analysis of pilot-operated regulators can be simplied by viewing
them as two independent regulators connected together. The
smaller of the two is generally the pilot.
Setpoint
We may think of the pilot as the “brains” of the system. Setpoint
and many performance variables are determined by the pilot. It
senses P2 directly and will continue to make changes in PL on
the main regulator until the system is in equilibrium. The main
regulator is the “muscle” of the system, and may be used to control
large ows and pressures.
Spring Action
Notice that the pilot uses a spring-open action as found in directoperated regulators. The main regulator, shown in Figure 1, uses
a spring-close action. The spring, rather than loading pressure, is
used to achieve shutoff. Increasing PL from the pilot onto the main
diaphragm opens the main regulator.
Pilot Advantage
Because the pilot is the controlling device, many of the
performance criteria we have discussed apply to the pilot. For
example, droop is determined mainly by the pilot. By using very
small pilot orices and light springs, droop can be made small.
Because of reduced droop, we will have greater usable capacity.
Pilot lockup determines the lockup characteristics for the system.
The main regulator spring provides tight shutoff whenever the pilot
is locked up.
Gain and Restrictions
Stability
Although increased gain (sensitivity) is often considered an
advantage, it also increases the gain of the entire pressure
regulator system. If the system gain is too high, it may become
unstable. In other words, the regulator might tend to oscillate;
over-reacting by continuously opening and closing. Pilot gain
can be modied to tune the regulator to the system. To provide
a means for changing gain, every pilot-operated regulator system
contains both a xed and a variable restriction. The relative size
of one restriction compared to the other can be varied to change
gain and speed of response.
Regulator Pilots
To improve the sensitivity of our regulator, we would like to be
able to sense P2 and then somehow make a change in loading
pressure (PL) that is greater than the change in P2. To accomplish
this, we can use a device called a pilot, or pressure amplier.
The major function of the pilot is to increase regulator sensitivity.
If we can sense a change in P2 and translate it into a larger change
in PL, our regulator will be more responsive (sensitive) to changes
in demand. In addition, we can signicantly reduce droop so its
effect on accuracy and capacity is minimized.
Gain
The amount of amplication supplied by the pilot is called
“gain”. To illustrate, a pilot with a gain of 20 will multiply
the effect of a 1 psi change on the main diaphragm by 20. For
example, a decrease in P2 opens the pilot to increase PL 20 times
as much.
Figure 1. Pilot-Operated Regulator
MAIN
REGULATOR
PILOT
REGULATOR
INLET PRESSURE, P
1
OUTLET PRESSURE, P
2
ATMOSPHERIC PRESSURE
LOADING PRESSURE, P
L
Page 27
Principles of Pilot-Operated Regulators
597
Te c h n i c a l
Loading and Unloading Designs
A loading pilot-operated design (Figure 2), also called two-path
control, is so named because the action of the pilot loads PL onto
the main regulator measuring element. The variable restriction, or
pilot orice, opens to increase PL.
An unloading pilot-operated design (Figure 3) is so named because
the action of the pilot unloads PL from the main regulator.
PILOT
FLOW
LARGER FIXED
RESTRICTION
VARIABLE
RESTRICTION
P
2
P
1
P
L
P
2
TO MAIN
REGULATOR
Figure 2. Fixed Restrictions and Gain (Used on Two-Path Control Systems)
Restrictions, Response Time, and Gain
Consider the example shown in Figure 2 with a small xed
restriction. Decreasing P2 will result in pressure PL increasing.
Increasing P2 will result in a decrease in PL while PL bleeds out
through the small xed restriction.
If a larger xed restriction is used with a variable restriction, the
gain (sensitivity) is reduced. A larger decrease in P2 is required to
increase PL to the desired level because of the larger xed restriction.
Two-Path Control (Loading Design)
In two-path control systems (Figure 4), the pilot is piped so that
P2 is registered on the pilot diaphragm and on the main regulator
diaphragm at the same time. When downstream demand is
constant, P2 positions the pilot diaphragm so that ow through the
pilot will keep P2 and PL on the main regulator diaphragm. When
P2 changes, the force on top of the main regulator diaphragm
and on the bottom of the pilot diaphragm changes. As P2 acts on
the main diaphragm, it begins repositioning the main valve plug.
This immediate reaction to changes in P2 tends to make two-path
designs faster than other pilot-operated regulators. Simultaneously,
P2 acting on the pilot diaphragm repositions the pilot valve and
Figure 3. Unloading Systems
Figure 4. Two-Path Control
VARIABLE
RESTRICTION
FLOW
FIXED RESTRICTION
PILOT
FLOW
P
2
P
1
P
L
VARIABLE
RESTRICTION
SMALLER FIXED
RESTRICTION
P
2
TO MAIN
REGULATOR
INLET PRESSURE, P
1
OUTLET PRESSURE, P
2
ATMOSPHERIC PRESSURE
LOADING PRESSURE, P
L
PILOT
FLOW
FIXED
RESTRICTION
VARIABLE
RESTRICTION
P
2
P
1
P
L
P
2
TO MAIN
REGULATOR
Page 28
Principles of Pilot-Operated Regulators
598
Te c h n i c a l
changes PL on the main regulator diaphragm. This adjustment to
PL accurately positions the main regulator valve plug. PL on the
main regulator diaphragm bleeds through a xed restriction until
the forces on both sides are in equilibrium. At that point, ow
through the regulator valve matches the downstream demand.
Two-Path Control Advantages
The primary advantages of two-path control are speed and
accuracy. These systems may limit droop to less than 1%. They
are well suited to systems with requirements for high accuracy,
large capacity, and a wide range of pressures.
Unloading Control
Unloading systems (Figure 5) locate the pilot so that P2 acts only
on the pilot diaphragm. P1 constantly loads under the regulator
diaphragm and has access to the top of the diaphragm through a
xed restriction.
When downstream demand is constant, the pilot valve is open
enough that PL holds the position of the main regulator diaphragm.
When downstream demand changes, P2 changes and the pilot
diaphragm reacts accordingly. The pilot valve adjusts PL to
reposition and hold the main regulator diaphragm.
Unloading Control Advantages
Unloading systems are not quite as fast as two-path systems,
and they can require higher differential pressures to operate.
However, they are simple and more economical, especially in
large regulators. Unloading control is used with popular elastomer
diaphragm style regulators. These regulators use a exible
membrane to throttle ow.
Performance Summary
Accuracy
Because of their high gain, pilot-operated regulators are extremely
accurate. Droop for a direct-operated regulator might be in the
range of 10 to 20 % whereas pilot-operated regulators are between
one and 3% with values under 1% possible.
Capacity
Pilot-operated designs provide high capacity for two reasons.
First, we have shown that capacity is related to droop. And
because droop can be made very small by using a pilot, capacity is
increased. In addition, the pilot becomes the “brains” of the system
and controls a larger, sometimes much larger, main regulator. This
also allows increased ow capabilities.
DOWNSTREAM PRESSURE (P
2
)
HIGH CAPACITY
FLOW RATE
MINIMAL
DROOP
Figure 6. Pilot-Operated Regulator Performance Figure 5. Unloading Control
MAIN REGULATOR
DIAPHRAGM
FIXED
RESTRICTION
INLET PRESSURE, P
1
OUTLET PRESSURE, P
2
ATMOSPHERIC PRESSURE
LOADING PRESSURE, P
L
Page 29
Principles of Pilot-Operated Regulators
599
Te c h n i c a l
Figure 8. Type 99, Typical Two-Path Control
with Integrally Mounted Pilot
Lockup
The lockup characteristics for a pilot-operated regulator are the
lockup characteristics of the pilot. Therefore, with small orices,
lockup pressures can be small.
Applications
Pilot-operated regulators should be considered whenever accuracy,
capacity, and/or high pressure are important selection criteria.
They can often be applied to high capacity services with greater
economy than a control valve and actuator with controller.
Two-Path Control
In some designs (Figure 7), the pilot and main regulator are
separate components. In others (Figure 8), the system is integrated
into a single package. All, however, follow the basic design
concepts discussed earlier.
Type 1098-EGR
The schematic in Figure 7 illustrates the Type 1098-EGR
regulator’s operation. It can be viewed as a model for all twopath, pilot-operated regulators. The pilot is simply a sensitive
direct-operated regulator used to send loading pressure to the main
regulator diaphragm.
Identify the inlet pressure (P1). Find the downstream pressure (P2).
Follow it to where it opposes the loading pressure on the main
regulator diaphragm. Then, trace P2 back to where it opposes the
control spring in the pilot. Finally, locate the route of P2 between
the pilot and the regulator diaphragm.
Changes in P2 register on the pilot and main regulator diaphragms
at the same time. As P2 acts on the main diaphragm, it begins
repositioning the main valve plug. Simultaneously, P2 acting on
the pilot diaphragm repositions the pilot valve and changes PL on
the main regulator diaphragm. This adjustment in PL accurately
positions the main regulator valve plug.
Figure 7. Type 1098-EGR, Typical Two-Path Control
INLET PRESSURE, P
1
OUTLET PRESSURE, P
2
ATMOSPHERIC PRESSURE
LOADING PRESSURE, P
L
INLET PRESSURE, P
1
OUTLET PRESSURE, P
2
ATMOSPHERIC PRESSURE
LOADING PRESSURE, P
L
Type 1098-EGR
A6563
A6469
Page 30
Principles of Pilot-Operated Regulators
600
Te c h n i c a l
As downstream demand is met, P2 rises. Because P2 acts directly
on both the pilot and main regulator diaphragms, this design
provides fast response.
Type 99
The schematic in Figure 8 illustrates another typical two-path
control design, the Type 99. The difference between the
Type 1098-EGR and the Type 99 is the integrally mounted pilot
of the Type 99.
The pilot diaphragm measures P2. When P2 falls below the pilot
setpoint, the diaphragm moves away from the pilot orice and
allows loading pressure to increase. This loads the top of the
main regulator diaphragm and strokes the main regulator valve
open further.
Unloading Design
Unloading designs incorporate a molded composition diaphragm
that serves as the combined loading and restricting component of
the main regulator. Full upstream pressure (P1) is used to load the
regulator diaphragm when it is seated. The regulator shown in
Figure 9 incorporates an elastomeric valve closure member.
Figure 9. Type EZR, Unloading Design
Unloading regulator designs are slower than two-path control
systems because the pilot must rst react to changes in P2 before
the main regulator valve moves. Recall that in two-path designs,
the pilot and main regulator diaphragms react simultaneously.
P1 passes through a xed restriction and lls the space above the
regulator diaphragm. This xed restriction can be adjusted to
increase or decrease regulator gain. P1 also lls the cavity below
the regulator diaphragm. Because the surface area on the top side
of the diaphragm is larger than the area exposed to P1 below, the
diaphragm is forced down against the cage to close the regulator.
When downstream demand increases, the pilot opens. When the
pilot opens, regulator loading pressure escapes downstream much
faster than P1 can bleed through the xed restriction. As pressure
above the regulator diaphragm decreases, P1 forces the diaphragm
away from its seat.
When downstream demand is reduced, P2 increases until it’s
high enough to compress the pilot spring and close the pilot
valve. As the pilot valve closes, P1 continues to pass through the
xed restriction and ows into the area above the main regulator
diaphragm. This loading pressure, PL, forces the diaphragm back
toward the cage, reducing ow through the regulator.
INLET PRESSURE, P
1
OUTLET PRESSURE, P
2
ATMOSPHERIC PRESSURE
LOADING PRESSURE, P
L
W7438
Page 31
Selecting and Sizing Pressure Reducing Regulators
601
Te c h n i c a l
Introduction
Those who are new to the regulator selection and sizing process
are often overwhelmed by the sheer number of regulator types
available and the seemingly endless lists of specications in
manufacturer’s literature. This application guide is designed
to assist you in selecting a regulator that ts your application’s
specic needs.
Although it might seem obvious, the rst step is to consider the
application itself. Some applications immediately point to a group
of regulators designed specically for that type of service. The
Application Guide has sections to help identify regulators that
are designed for specic applications. There are Application
Maps, Quick Selection Guides, an Applications section, and
Product Pages. The Application Map shows some of the common
applications and the regulators that are used in those applications.
The Quick Selection Guide lists the regulators by application, and
provides important selection information about each regulator.
The Applications section explains the applications covered in the
section and it also explains many of the application considerations.
The Product Pages provide specic details about the regulators
that are suitable for the applications covered in the section. To
begin selecting a regulator, turn to the Quick Selection Guide in the
appropriate Applications section.
Quick Selection Guides
Quick Selection Guides identify the regulators with the appropriate
pressure ratings, outlet pressure ranges, and capacities. These
guides quickly narrow the range of potentially appropriate
regulators. The choices identied by using a Quick Selection
Guide can be narrowed further by using the Product Pages to nd
more information about each of the regulators.
Product Pages
Identifying the regulators that can pass the required ow narrows
the possible choices further. When evaluating ow requirements,
consider the minimum inlet pressure and maximum ow
requirements. Again, this worst case combination ensures that the
regulator can pass the required ow under all anticipated conditions.
After one or more regulators have been identied as potentially
suitable for the service conditions, consult specic Product Pages
to check regulator specications and capabilities. The application
requirements are compared to regulator specications to narrow the
range of appropriate selections. The following specications can
be evaluated in the Product Pages:
• Product description and available sizes
• Maximum inlet and outlet pressures (operating
and emergency)
• Outlet pressure ranges
• Flow capacity
• End connection styles
• Regulator construction materials
• Accuracy
• Pressure registration (internal or external)
• Temperature capabilities
After comparing the regulator capabilities with the application
requirements, the choices can be narrowed to one or a few regulators.
Final selection might depend upon other factors including special
requirements, availability, price, and individual preference.
Special Requirements
Finally, evaluate any special considerations, such as the need for
external control lines, special construction materials, or internal
overpressure protection. Although overpressure protection might
be considered during sizing and selection, it is not covered in
this section.
The Role of Experience
Experience in the form of knowing what has worked in the
past, and familiarity with specic products, has great value in
regulator sizing and selection. Knowing the regulator performance
characteristics required for a specic application simplies the
process. For example, when fast speed of response is required, a
direct-operated regulator may come to mind; or a pilot-operated
regulator with an auxiliary, large capacity pilot to speed changes in
loading pressure.
Sizing Equations
Sizing equations are useful when sizing pilot-operated regulators
and relief valves. They can also be used to calculate the wide-
open ow of direct-operated regulators. Use the capacity tables
or curves in this application guide when sizing direct-operated
regulators and relief/backpressure regulators. The sizing equations
are in the Valve Sizing Calculations section.
Page 32
Selecting and Sizing Pressure Reducing Regulators
602
Te c h n i c a l
General Sizing Guidelines
The following are intended to serve only as guidelines when sizing
pressure reducing regulators. When sizing any regulator, consult
with experienced personnel or the regulator manufacturer for
additional guidance and information relating to specic applications.
Body Size
Regulator body size should never be larger than the pipe size.
However, a properly sized regulator may be smaller than the pipeline.
Construction
Be certain that the regulator is available in materials that are
compatible with the controlled uid and the temperatures used. Also,
be sure that the regulator is available with the desired end connections.
Pressure Ratings
While regulators are sized using minimum inlet pressures to
ensure that they can provide full capacity under all conditions, pay
particular attention to the maximum inlet and outlet pressure ratings.
Wide-Open Flow Rate
The capacity of a regulator when it has failed wide-open is usually
greater than the regulating capacity. For that reason, use the
regulating capacities when sizing regulators, and the wide-open
ow rates only when sizing relief valves.
Outlet Pressure Ranges and Springs
If two or more available springs have published outlet pressure
ranges that include the desired pressure setting, use the spring
with the lower range for better accuracy. Also, it is not necessary
to attempt to stay in the middle of a spring range, it is acceptable
to use the full published outlet pressure range without sacricing
spring performance or life.
Accuracy
Of course, the need for accuracy must be evaluated. Accuracy is
generally expressed as droop, or the reduction of outlet pressure
experienced as the ow rate increases. It is stated in percent,
inches of water column, or pounds per square inch. It indicates
the difference between the outlet pressure at low ow rates and the
outlet pressure at the published maximum ow rate. Droop is also
called offset or proportional band.
Inlet Pressure Losses
The regulator inlet pressure used for sizing should be measured
directly at the regulator inlet. Measurements made at any distance
upstream from the regulator are suspect because line loss can
signicantly reduce the actual inlet pressure to the regulator. If
the regulator inlet pressure is given as a system pressure upstream,
some compensation should be considered. Also, remember that
downstream pressure always changes to some extent when inlet
pressure changes.
Orifice Diameter
The recommended selection for orice size is the smallest diameter
that will handle the ow. This can benet operation in several
ways: instability and premature wear might be avoided, relief
valves may be smaller, and lockup pressures may be reduced.
Speed of Response
Direct-operated regulators generally have faster response to quick
ow changes than pilot-operated regulators.
Turn-Down Ratio
Within reasonable limits, most soft-seated regulators can maintain
pressure down to zero ow. Therefore, a regulator sized for a high
ow rate will usually have a turndown ratio sufcient to handle
pilot-light sized loads during periods of low demand.
Sizing Exercise: Industrial Plant Gas Supply
Regulator selection and sizing generally requires some subjective
evaluation and decision making. For those with little experience,
the best way to learn is through example. Therefore, these
exercises present selection and sizing problems for practicing the
process of identifying suitable regulators.
Our task is to select a regulator to supply reduced pressure natural
gas to meet the needs of a small industrial plant. The regulated gas
is metered before entering the plant. The selection parameters are:
• Minimum inlet pressure, P
1min
= 30 psig
• Maximum inlet pressure, P
1max
= 40 psig
• Outlet pressure setting, P
2
= 1 psig
• Flow, Q = 95 000 SCFH
• Accuracy (droop required) = 10% or less
Page 33
Selecting and Sizing Pressure Reducing Regulators
603
Te c h n i c a l
Quick Selection Guide
Turn to the Commercial/Industrial Quick Selection Guide. From
the Quick Selection Guide, we nd that the choices are:
• Type 133
• Type 1098-EGR
Product Pages
Under the product number on the Quick Selection Guide is the
page number of the product page. Look at the ow capacities of
each of the possible choices. From the product pages we found
the following:
Q = 95 000 SCFH
P2 = 1 PSIG
GAS SUPPLY = 30 TO 40 PSIG
M
METER
INDUSTRIAL PLANT
Figure 1. Natural Gas Supply
• At 30 psig inlet pressure and 10% droop, the Type 133 has a
ow capacity of 90 000 SCFH. This regulator does not meet
the required ow capacity.
• At 30 psig inlet pressure, the Type 1098-EGR has a ow
capacity of 131 000 SCFH. By looking at the Proportional
Band (Droop) table, we see that the Type 6352 pilot with the
yellow pilot spring and the green main valve has 0.05 psig
droop. This regulator meets the selection criteria.
Final Selection
We nd that the Type 1098-EGR meets the selection criteria.
Page 34
Overpressure Protection Methods
604
Te c h n i c a l
Overpressure protective devices are of vital concern. Safety codes
and current laws require their installation each time a pressure
reducing station is installed that supplies gas from any system to
another system with a lower maximum allowable operating pressure.
Methods of Overpressure Protection
The most commonly used methods of overpressure protection, not
necessarily in order of use or importance, include:
• Relief Valves (Figure 1)
• Monitors (Figures 2 and 3)
• Series Regulation (Figure 4)
• Shutoff (Figure 5)
• Relief Monitor (Figure 6)
Figure 1. Relief Valve Schematic
Relief Valves
A relief valve is a device that vents process uid to atmosphere to
maintain the pressure downstream of the regulator below the safe
maximum pressure. Relief is a common form of overpressure
protection typically used for low to medium capacity applications.
(Note: Fisher® relief valves are not ASME safety relief valves.)
Types of Relief Valves
The basic types of relief valves are:
• Pop type
• Direct-operated relief valves
• Pilot-operated relief valves
• Internal relief valves
The pop type relief valve is the simplest form of relief. Pop relief
valves tend to go wide-open once the pressure has exceeded its
setpoint by a small margin. The setpoint can drift over time,
and because of its quick opening characteristic the pop relief can
sometimes become unstable when relieving, slamming open and
closed. Many have a non-adjustable setpoint that is set and pinned
at the factory.
If more accuracy is required from a relief valve, the direct-operated
relief valve would be the next choice. They can throttle better
than a pop relief valve, and tend to be more stable, yet are still
relatively simple. Although there is less drift in the setpoint of the
direct-operated relief valve, a signicant amount of build-up is
often required to obtain the required capacity.
The pilot-operated relief valves have the most accuracy, but are
also the most complicated and expensive type of relief. They use a
pilot to dump loading pressure, fully stroking the main valve with
very little build-up above setpoint. They have a large capacity and
are available in larger sizes than other types of relief.
Many times, internal relief will provide adequate protection for a
downstream system. Internal relief uses a relief valve built into the
regulator for protection. If the pressure builds too far above the
setpoint of the regulator, the relief valve in the regulator opens up,
allowing excess pressure to escape through the regulator vent.
Advantages
The relief valve is considered to be the most reliable type of
overpressure protection because it is not subject to blockage
by foreign objects in the line during normal operations. It
also imposes no decrease in the regulator capacity which it is
protecting, and it has the added advantage of being its own
alarm when it vents. It is normally reasonable in cost and keeps
the customer in service despite the malfunction of the pressure
reducing valve.
Disadvantages
When the relief valve blows, it could possibly create a hazard in
the surrounding area by venting. The relief valve must be sized
carefully to relieve the gas or uid that could ow through the
pressure reducing valve at its maximum inlet pressure and in
the wide-open position, assuming no ow to the downstream.
Therefore, each application must be sized individually. The
requirement for periodic testing of relief valves also creates an
operational and/or public relations problem.
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Monitoring Regulators
Monitoring is overpressure control by containment. When the
working pressure reducing valve ceases to control the pressure,
a second regulator installed in series, which has been sensing
the downstream pressure, goes into operation to maintain the
downstream pressure at a slightly higher than normal pressure.
The monitoring concept is gaining in popularity, especially in
low-pressure systems, because very accurate relay pilots permit
reasonably close settings of the working and monitoring regulators.
The two types of wide-open monitoring are upstream and
downstream monitoring. One question often asked is, “Which
is better, upstream or downstream monitoring?” Using two
identical regulators, there is no difference in overall capacity with
either method.
When using monitors to protect a system or customer who may at
times have zero load, a small relief valve is sometimes installed
downstream of the monitor system with a setpoint just above the
monitor. This allows for a token relief in case dust or dirt in the
system prevents bubble tight shutoff of the regulators.
Advantages
The major advantage is that there is no venting to atmosphere.
During an overpressure situation, monitoring keeps the customer
on line and keeps the downstream pressure relatively close to the
setpoint of the working regulator. Testing is relatively easy and
safe. To perform a periodic test on a monitor, increase the outlet
set pressure of the working device and watch the pressure to
determine if the monitor takes over.
Disadvantages
Compared to relief valves, monitoring generally requires a higher
initial investment. Monitoring regulators are subject to blocking,
which is why lters or strainers are specied with increasing
frequency. Because the monitor is in series, it is an added
restriction in the line. This extra restriction can sometimes force
one to use a larger, more expensive working regulator.
Figure 2. Monitoring Regulators Schematic
Working Monitor
A variation of monitoring overpressure protection that overcomes
some of the disadvantages of a wide-open monitor is the “working
monitor” concept wherein a regulator upstream of the working
regulator uses two pilots. This additional pilot permits the
monitoring regulator to act as a series regulator to control an
intermediate pressure during normal operation. In this way, both
units are always operating and can be easily checked for proper
operation. Should the downstream pressure regulator fail to
control, however, the monitoring pilot takes over the control at a
slightly higher than normal pressure and keeps the customer on
line. This is pressure control by containment and eliminates public
relations problems.
Figure 3. Working Monitor Schematic
Figure 4. Series Regulation Schematic
Series Regulation
Series regulation is also overpressure protection by containment
in that two regulators are set in the same pipeline. The rst unit
maintains an inlet pressure to the second valve that is within the
maximum allowable operating pressure of the downstream system.
Under this setup, if either regulator should fail, the resulting
downstream pressure maintained by the other regulator would not
exceed the safe maximum pressure.
This type of protection is normally used where the regulator station
is reducing gas to a pressure substantially below the maximum
allowable operating pressure of the distribution system being
supplied. Series regulation is also found frequently in farm taps
and in similar situations within the guidelines mentioned above.
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Te c h n i c a l
Advantages
Again, nothing is vented to atmosphere.
Disadvantages
Because the intermediate pressure must be cut down to a pressure
that is safe for the entire downstream, the second-stage regulator
often has very little pressure differential available to create ow.
This can sometimes make it necessary to increase the size of the
second regulator signicantly. Another drawback occurs when the
rst-stage regulator fails and no change in the nal downstream
pressure is noticed because the system operates in what appears
to be a “normal” manner without benet of protection. Also, the
rst-stage regulator and intermediate piping must be capable of
withstanding and containing maximum upstream pressure.
The second-stage regulator must also be capable of handling the
full inlet pressure in case the rst-stage unit fails to operate. In
case the second-stage regulator fails, its actuator will be subjected
to the intermediate pressure set by the rst-stage unit. The secondstage actuator pressure ratings should reect this possibility.
Advantages
By shutting off the customer completely, the safety of the
downstream system is assured. Again, there is no public relations
problem or hazard from venting gas or other media.
Disadvantages
The customer may be shutoff because debris has temporarily
lodged under the seat of the operating regulator, preventing tight
shutoff. A small relief valve can take care of this situation.
On a distribution system with a single supply, using a slam-shut
can require two trips to each customer, the rst to shutoff the
service valve, and the second visit after the system pressure has
been restored to turn the service valve back on and re-light the
appliances. In the event a shutoff is employed on a service line
supplying a customer with processes such as baking, melting
metals, or glass making, the potential economic loss could dictate
the use of an overpressure protection device that would keep the
customer online.
Another problem associated with shutoffs is encountered when the
gas warms up under no-load conditions. For instance, a regulator
locked up at approximately 7-inches w.c. could experience a
pressure rise of approximately 0.8-inch w.c. per degree Fahrenheit
rise, which could cause the high-pressure shutoff to trip when there
is actually no equipment failure.
Figure 5. Shutoff Schematic
Figure 6. Relief Monitor Schematic
Shutoff Devices
The shutoff device also accomplishes overpressure protection
by containment. In this case, the customer is shutoff completely
until the cause of the malfunction is determined and the device is
manually reset. Many gas distribution companies use this as an
added measure of protection for places of public assembly such as
schools, hospitals, churches, and shopping centers. In those cases,
the shutoff device is a secondary form of overpressure protection.
Shutoff valves are also commonly used by boiler manufacturers in
combustion systems.
Relief Monitor
Another concept in overpressure protection for small industrial
and commercial loads, up to approximately 10 000 cubic feet per
hour, incorporates both an internal relief valve and a monitor. In
this device, the relief capacity is purposely restricted to prevent
excess venting of gas in order to bring the monitor into operation
more quickly. The net result is that the downstream pressure is
protected, in some cases to less than 1 psig. The amount of gas
vented under maximum inlet pressure conditions does not exceed
the amount vented by a domestic relief type service regulator.
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With this concept, the limitation by regulator manufacturers of
inlet pressure by orice size, as is found in “full relief” devices,
is overcome. Downstream protection is maintained, even with
abnormally high inlet pressure. Public relations problems are
kept to a minimum by the small amount of vented gas. Also,
the unit does not require manual resetting, but can go back into
operation automatically.
Dust or dirt can clear itself off the seat, but if the obstruction to the
disk closing still exists when the load goes on, the customer would
be kept online. When the load goes off, the downstream pressure
will again be protected. During normal operation, the monitoring
portion of the relief monitor is designed to move slightly with
minor uctuations in downstream pressure or ow.
Summary
From the foregoing discussion, it becomes obvious that there are
many design philosophies available and many choices of equipment
to meet overpressure protection requirements. Also, assume the
overpressure device will be called upon to operate sometime after
it is installed. The overall design must include an analysis of the
conditions created when the protection device operates.
The accompanying table shows:
• What happens when the various types of overpressure
protection devices operate
• The type of reaction required
• The effect upon the customer or the public
• Some technical conditions
These are the general characteristics of the various types of safety
devices. From the conditions and results shown, it is easier to
decide which type of overpressure equipment best meets your
needs. Undoubtedly, compromises will have to be made between
the conditions shown here and any others which may govern your
operating parameters.
Types of Overpressure Protection
PERFORMANCE QUESTIONS RELIEF
WORKING
MONITOR
MONITOR
SERIES
REGULATION
SHUTOFF
RELIEF
MONITOR
Keeps application online? Yes Yes Yes Yes No Yes
Venting to atmosphere? Yes No No No No Minor
Manual resetting required after operation? No No No No Yes No
Reduces capacity of regulator? No Yes Yes Yes No No
Constantly working during normal operation? No Yes No Yes No Yes
Demands “emergency” action? Yes No No No Yes Maybe
Will surveillance of pressure charts indicate partial loss of
performance of overpressure devices?
No Yes Maybe Yes No No
Will surveillance of pressure charts indicate regulator has failed and
safety device is in control?
Yes Yes Yes Yes Yes Yes
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Overpressure Protection
Overpressure protection is a primary consideration in the design
of any piping system. The objective of overpressure protection
is to maintain the pressure downstream of a regulator at a safe
maximum value.
Main Regulator
Pressure reducing regulators have different pressure ratings
which refer to the inlet, outlet, and internal components. The
lowest of these should be used when determining the maximum
allowable pressure.
Piping
Piping is limited in its ability to contain pressure. In addition to
any physical limitations, some applications must also conform to
one or more applicable pressure rating codes or regulations.
Relief Valves
Relief involves maintaining the pressure downstream of a regulator
at a safe maximum pressure using any device that vents uid to
a lower pressure system (often the atmosphere). Relief valve
exhaust must be directed or piped to a safe location. Relief valves
perform this function. They are considered to be one of the most
reliable types of overpressure protection available and are available
in a number of different types. Fisher® relief valves are not ASME
safety relief valves.
NATURAL GAS
TRANSMISSION LINE
REGULATOR
RELIEF VALVE
RELIEF VALVEREGULATOR
TYPE 289 DIRECT-OPERATED
TYPE 627 WITH INTERNAL RELIEF
TYPE 289P-6358 PILOT-OPERATED
H200 SERIES POP TYPE
In the system shown in Figure 1, a high-pressure transmission
system delivers natural gas through a pressure reducing regulator
to a lower pressure system that distributes gas to individual
customers. The regulators, the piping, and the devices that
consume gas are protected from overpressure by relief valves.
The relief valve’s setpoint is adjusted to a level established by
the lowest maximum pressure rating of any of the lower pressure
system components.
Maximum Pressure Considerations
Overpressure occurs when the pressure of a system is above the
setpoint of the device controlling its pressure. It is evidence of
some failure in the system (often the upstream regulator), and it
can cause the entire system to fail if it’s not limited. To implement
overpressure protection, the weakest part in the pressure system
is identied and measures are taken to limit overpressure to that
component’s maximum pressure rating. The most vulnerable
components are identied by examining the maximum pressure
ratings of the:
• Downstream equipment
• Low-pressure side of the main regulator
• Piping
The lowest maximum pressure rating of the three is the maximum
allowable pressure.
Downstream Equipment
The downstream component (appliance, burner, boiler, etc.) with
the lowest maximum pressure rating sets the highest pressure that
all the downstream equipment can be subjected to.
Figure 2. Types of Relief Valves
Figure 1. Distribution System
P
P
M
W1921
W3167
W4793
W1870
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Principles of Relief Valves
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Te c h n i c a l
Relief Valve Popularity
Relief valves are popular for several reasons. They do not block
the normal ow through a line. They do not decrease the capacities
of the regulators they protect. And, they have the added advantage
of being an alarm if they vent to atmosphere.
Relief Valve Types
Relief valves are available in four general types. These include:
pop type, direct-operated, pilot-operated, and internal relief valves.
Selection Criteria
Pressure Build-up
Cost Versus Performance
Given several types of relief valves to choose from, selecting one
type is generally based on the ability of the valve to provide adequate
protection at the most economical cost. Reduced pressure build-up
and increased capacity generally come at an increased price.
Installation and Maintenance Considerations
Initial costs are only a part of the overall cost of ownership.
Maintenance and installation costs must also be considered over
the life of the relief valve. For example, internal relief might be
initially more economical than an external relief valve. However,
maintaining a regulator with internal relief requires that the
system be shut down and the regulator isolated. This may involve
additional time and the installation of parallel regulators and relief
valves if ow is to be maintained to the downstream system during
maintenance operations.
PRESSURE BUILD-UP
RELIEF VALVE
SETPOINT
FLOW
PRESSURE
Figure 3. Pressure Build-up
A relief valve has a setpoint at which it begins to open. For the
valve to fully open and pass the maximum ow, pressure must
build up to some level above the setpoint of the relief valve. This
is known as pressure build-up over setpoint, or simply build-up.
Periodic Maintenance
A relief valve installed in a system that normally performs
within design limits is very seldom exercised. The relief valve
sits and waits for a failure. If it sits for long periods it may not
perform as expected. Disks may stick in seats, setpoints can
shift over time, and small passages can become clogged with
pipeline debris. Therefore, periodic maintenance and inspection
is recommended. Maintenance requirements might inuence
the selection of a relief valve.
Figure 4. Pop Type Relief Valve Construction and Operation
LOADING SPRING
POPPET
SOFT DISK
SEAT RING
CLOSED CLOSED WIDE-OPEN
Pop Type Relief Valve
The most simple type of relief valve is the pop type. They are used
wherever economy is the primary concern and some setpoint drift
is acceptable.
Operation
Pop type relief valves are essentially on-off devices. They
operate in either the closed or wide-open position. Pop type
designs register pressure directly on a spring-opposed poppet.
The poppet assembly includes a soft disk for tight shutoff against
the seat ring. When the inlet pressure increases above setpoint,
the poppet assembly is pushed away from the seat. As the
poppet rises, pressure registers against a greater surface area of
the poppet. This dramatically increases the force on the poppet.
Therefore, the poppet tends to travel to the fully open position
reducing pressure build-up.
W0121
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Build-up Over Setpoint
Recall that pressure build-up relates capacity to pressure;
increasing capacity requires some increase in pressure. In
throttling relief valves, pressure build-up is related to accuracy.
In pop type relief valves, build-up over setpoint results largely
because the device is a restriction to ow rather than the spring
rate of the valve’s loading spring.
Fixed Setpoint
The setpoint of a pop type valve cannot be adjusted by the user.
The spring is initially loaded by the manufacturer. A pinned spring
retainer keeps the spring in position. This is a safety measure that
prevents tampering with the relief valve setpoint.
Typical Applications
This type of relief valve may be used where venting to the
atmosphere is acceptable, when the process uid is compatible
with the soft disk, and when relief pressure variations are
allowable. They are often used as inexpensive token relief. For
example, they may be used simply to provide an audible signal of
an overpressure condition.
These relief valves may be used to protect against overpressure
stemming from a regulator with a minimal amount of seat leakage.
Unchecked, this seat leakage could allow downstream pressure to
build to full P1 over time. The use of a small pop type valve can
be installed to protect against this situation.
These relief valves are also commonly installed with a regulator in
a natural gas system farm tap, in pneumatic lines used to operate
air drills, jackhammers, and other pneumatic equipment, and in
many other applications.
Advantages
Pop type relief valves use few parts. Their small size allows
installation where space is limited. Also, low initial cost, easy
installation, and high capacity per dollar invested can result in
economical system relief.
Disadvantages
The setpoint of a pop type relief valve may change over time.
The soft disk may stick to the seat ring and cause the pop
pressure to increase.
As an on-off device, this style of relief valve does not throttle ow
over a pressure range. Because of its on-off nature, this type of
relief valve may create pressure surges in the downstream system.
Figure 5. Direct-Operated Relief Valve Schematic
If the relief valve capacity is signicantly larger than the failed
regulator’s capacity, the relief valve may over-compensate each
time it opens and closes. This can cause the downstream pressure
system to become unstable and cycle. Cycling can damage the
relief valve and downstream equipment.
Direct-Operated Relief Valves
Compared to pop type relief valves, direct-operated relief valves
provide throttling action and may require less pressure build-up to
open the relief valve.
Operation
A schematic of a direct-operated relief valve is shown in Figure 5.
It looks like an ordinary direct-operated regulator except that it
senses upstream pressure rather than downstream pressure. And,
it uses a spring-close rather than a spring-open action. It contains
the same essential elements as a direct-operated regulator:
• A diaphragm that measures system pressure
• A spring that provides the initial load to the diaphragm and is
used to establish the relief setpoint
• A valve that throttles the relief ow
DIAPHRAGM
SPRING
VALVE
VENT
SYSTEM PRESSURE
LOWER-PRESSURE SYSTEM
(USUALLY ATMOSPHERE)
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Principles of Relief Valves
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Te c h n i c a l
M1048
Type 289H
FLOW
CONTROL
LINE
PILOT
RESTRICTION
MAIN RELIEF
VALVE
PLUG AND
SEAT RING
PILOT
EXHAUST
PLUG AND SEAT RING MAIN VALVE
Opening the Valve
As the inlet pressure rises above the setpoint of the relief valve, the
diaphragm is pushed upward moving the valve plug away from the
seat. This allows uid to escape.
Pressure Build-up Over Setpoint
As system pressure increases, the relief valve opens wider. This
allows more uid to escape and protects the system. The increase
in pressure above the relief setpoint that is required to produce
more ow through the relief valve is referred to as pressure
build-up. The spring rate and orice size inuence the amount of
pressure build-up that is required to fully stroke the valve.
Selection Criteria
Pressure Build-up
Some direct-operated relief valves require signicant pressure
build-up to achieve maximum capacity. Others, such as those
using pitot tubes, often pass high ow rates with minimal pressure
build-up. Direct-operated relief valves can provide good accuracy
within their design capacities.
Figure 6. Type 289 Relief Valve with Pitot Tube
Product Example
Pitot Tube
The relief valve shown in Figure 6 includes a pitot tube to reduce
pressure build-up. When the valve is opening, high uid velocity
through the seat ring creates an area of relatively low pressure.
Low pressure near the end of the pitot tube draws uid out of the
volume above the relief valve diaphragm and creates a partial
vacuum which helps to open the valve. The partial vacuum above
the diaphragm increases the relief valve capacity with less pressure
build-up over setpoint.
Typical Applications
Direct-operated relief valves are commonly used in natural gas
systems supplying commercial enterprises such as restaurants
and laundries, and in industry to protect industrial furnaces and
other equipment.
MAIN REGULATOR
EXHAUST
ELASTOMERIC
ELEMENT
ELASTOMERIC ELEMENT MAIN VALVE
Figure 7. Pilot-Operated Designs
CONTROL LINE
PILOT
RESTRICTION
MAIN RELIEF
VALVE
PILOT
EXHAUST
INLET PRESSURE
LOADING PRESSURE
EXHAUST
MAIN
REGULATOR EXHAUST
FLOW
ATMOSPHERIC
PRESSURE
LOADING SPRING
PITOT TUBE
DIAPHRAGM
VALVE DISK
SEAT RING
M1048
INLET PRESSURE
OUTLET PRESSURE
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Cost Versus Performance
The purchase price of a direct-operated relief valve is typically
lower than that of a pilot-operated design of the same size.
However, pilot-operated designs may cost less per unit of capacity
at very high ow rates.
Pilot-Operated Relief Valves
Pilot-operated relief valves utilize a pair of direct-operated relief
valves; a pilot and a main relief valve. The pilot increases the
effect of changes in inlet pressure on the main relief valve.
Operation
The operation of a pilot-operated relief valve is quite similar to
the operation of a pilot-operated pressure reducing regulator. In
normal operation, when system pressure is below setpoint of the
relief valve, the pilot remains closed. This allows loading pressure
to register on top of the main relief valve diaphragm. Loading
pressure on top of the diaphragm is opposed by an equal pressure
(inlet pressure) on the bottom side of the diaphragm. With little or
no pressure differential across the diaphragm, the spring keeps the
valve seated. Notice that a light-rate spring may be used because it
does not oppose a large pressure differential across the diaphragm.
The light-rate spring enables the main valve to travel to the wideopen position with little pressure build-up.
Increasing Inlet Pressure
When the inlet pressure rises above the relief setpoint, the pilot
spring is compressed and the pilot valve opens. The open pilot
bleeds uid out of the main valve spring case, decreasing pressure
above the main relief valve diaphragm. If loading pressure escapes
faster than it can be replaced through the restriction, the loading
pressure above the main relief valve diaphragm is reduced and the
relief valve opens. System overpressure exhausts through the vent.
Decreasing Inlet Pressure
If inlet pressure drops back to the relief valve setpoint, the pilot
loading spring pushes the pilot valve plug back against the
pilot valve seat. Inlet pressure again loads the main relief valve
diaphragm and closes the main valve.
Control Line
The control line connects the pilot with the pressure that is to be
limited. When overpressure control accuracy is a high priority, the
control line tap is installed where protection is most critical.
Product Example
Physical Description
This relief valve is a direct-operated relief valve with a pilot attached
(Figure 8). The pilot is a modied direct-operated relief valve, the
inlet pressure loads the diaphragm and ows through a restriction to
supply loading pressure to the main relief valve diaphragm.
Operation
During normal operation, the pilot is closed allowing loading
pressure to register above the main relief valve’s diaphragm.
This pressure is opposed by inlet pressure acting on the bottom
of the diaphragm.
If inlet pressure rises above setpoint, the pilot valve opens,
exhausting the loading pressure. If loading pressure is reduced
above the main relief valve diaphragm faster than it is replaced
through the pilot xed restriction, loading pressure is reduced and
inlet pressure below the diaphragm will cause the main regulator
to open.
Figure 8. Pilot-Operated Relief Valve
PILOT
MAIN VALVE
DIAPHRAGM
INLET PRESSURE
LOADING PRESSURE
EXHAUST
ATMOSPHERIC PRESSURE
E0061
TYPE 289P-6358B
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If inlet pressure falls below the relief set pressure, the pilot
spring will again close the pilot exhaust, increasing loading
pressure above the main relief valve diaphragm. This
increasing loading pressure causes the main valve to travel
towards the closed position.
Performance
Pilot-operated relief valves are able to pass large ow rates with a
minimum pressure build-up.
Typical Applications
Pilot-operated relief valves are used in applications requiring high
capacity and low pressure build-up.
Selection Criteria
Minimal Build-up
The use of a pilot to load and unload the main diaphragm and the
light-rate spring enables the main valve to travel wide-open with
little pressure build-up over setpoint.
Throttling Action
The sensitive pilot produces smooth throttling action when inlet
pressure rises above setpoint. This helps to maintain a steady
downstream system pressure.
Internal Relief
Regulators that include internal relief valves may eliminate the
requirement for external overpressure protection.
Operation
The regulator shown in Figure 9 includes an internal relief valve.
The relief valve has a measuring element (the main regulator
diaphragm), a loading element (a light spring), and a restricting
element (a valve seat and disk). The relief valve assembly is
located in the center of the regulator diaphragm.
Build-up Over Setpoint
Like other spring-loaded designs, internal relief valves will only
open wider if the inlet pressure increases. The magnitude of
pressure build-up is determined by the spring rates of the loading
spring plus the main spring. Both springs are considered because
they act together to resist diaphragm movement when pressure
exceeds the relief valve setpoint.
Product Example
A typical internal relief regulator construction is shown in
Figure 9. The illustration on the left shows the regulator with
both the relief valve and regulator valve in the closed position.
The illustration on the right shows the same unit after the inlet
Figure 9. Internal Relief Design
REGULATORS THAT INCLUDE INTERNAL RELIEF VALVES OFTEN ELIMINATE THE REQUIREMENT FOR EXTERNAL OVERPRESSURE PROTECTION. THE ILLUSTRATION ON THE LEFT SHOWS THE REGULATOR
WITH BOTH THE RELIEF VALVE AND THE REGULATOR VALVE IN THE CLOSED POSITION. THE ILLUSTRATION ON THE RIGHT SHOWS THE SAME UNIT AFTER P2 HAS INCREASED ABOVE THE
RELIEF VALVE SETPOINT. THE DIAPHRAGM HAS MOVED OFF THE RELIEF VALVE SEAT ALLOWING FLOW (EXCESS PRESSURE) TO EXHAUST THROUGH THE SCREENED VENT.
RELIEF VALVE CLOSED
MAIN VALVE CLOSED
MAIN VALVE CLOSED
INLET PRESSURE
RELIEF PRESSURE
ATMOSPHERIC PRESSURE
INLET PRESSURE
RELIEF PRESSURE
ATMOSPHERIC PRESSURE
LARGE OPENING FOR RELIEF
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pressure has increased above the relief valve setpoint. The
diaphragm has moved off the relief valve seat allowing the excess
pressure to exhaust through the vent.
Performance and Typical Applications
This design is available in congurations that can protect many
pressure ranges and ow rates. Internal relief is often used in
applications such as farm taps, industrial applications where
atmospheric exhaust is acceptable, and house service regulators.
Selection Criteria
Pressure Build-up
Relief setpoint is determined by a combination of the relief valve
and regulator springs; this design generally requires signicant
pressure build-up to reach its maximum relief ow rate. For the
same reason, internal relief valves have limited relief capacities.
They may provide full relief capacity, but should be carefully sized
for each application.
Space
Internal relief has a distinct advantage when there is not enough
space for an external relief valve.
Cost versus Performance
Because a limited number of parts are simply added to the regulator,
this type of overpressure protection is relatively inexpensive
compared to external relief valves of comparable capacity.
Maintenance
Because the relief valve is an integral part of the regulator’s
diaphragm, the regulator must be taken out of service when
maintenance is performed. Therefore, the application should be
able to tolerate either the inconvenience of intermittent supply or
the expense of parallel regulators and relief valves.
Selection and Sizing Criteria
There are a number of common steps in the relief valve
selection and sizing process. For every application, the
maximum pressure conditions, the wide-open regulator
ow capacity, and constant downstream demand should be
determined. Finally, this information is used to select an
appropriate relief valve for the application.
Maximum Allowable Pressure
Downstream equipment includes all the components of the system
that contain pressure; household appliances, tanks, tools, machines,
outlet rating of the upstream regulators, or other equipment. The
component with the lowest maximum pressure rating establishes
the maximum allowable system pressure.
Regulator Ratings
Pressure reducing regulators upstream of the relief valve have
ratings for their inlet, outlet, and internal components. The
lowest rating should be used when determining maximum
allowable pressure.
Piping
Piping pressure limitations imposed by governmental agencies,
industry standards, manufacturers, or company standards should be
veried before dening the maximum overpressure level.
Maximum Allowable System Pressure
The smallest of the pressure ratings mentioned above should be
used as the maximum allowable pressure. This pressure level
should not be confused with the relief valve setpoint which must be
set below the maximum allowable system pressure.
Determining Required Relief Valve Flow
A relief valve must be selected to exhaust enough ow to
prevent the pressure from exceeding the maximum allowable
system pressure. To determine this ow, review all upstream
components for the maximum possible ow that will cause
overpressure. If overpressure is caused by a pressure reducing
regulator, use the regulator’s wide-open ow coefcient to
calculate the required ow of the relief valve. This regulator ’s
wide-open ow is larger than the regulating ow used to select
the pressure reducing regulator.
Sizing equations have been developed to standardize valve
sizing. Refer to the Valve Sizing Calculations section to nd these
equations and explanations on how they are used.
Page 45
Principles of Relief Valves
615
Te c h n i c a l
Determine Constant Demand
In some applications, the required relief capacity can be reduced by
subtracting any load that is always on the system. This procedure
should be approached with caution because it may be difcult to
predict the worst-case scenario for downstream equipment failures.
It may also be important to compare the chances of making a
mistake in predicting the level of continuous ow consumption
with the potential negative aspects of an error. Because of the
hazards involved, relief valves are often sized assuming no
continuous ow to downstream equipment.
Selecting Relief Valves
Required Information
We have already reviewed the variables required to calculate the
regulator’s wide-open ow rate. In addition, we need to know
the type and temperature of the uid in the system, and the size of
the piping. Finally, if a vent stack will be required, any additional
build-up due to vent stack resistance should be considered.
Regulator Lockup Pressure
A relief valve setpoint is adjusted to a level higher than the
regulator’s lockup pressure. If the relief valve setpoint overlaps
lockup pressure of the regulator, the relief valve may open while
the regulator is still attempting to control the system pressure.
Identify Appropriate Relief Valves
Once the size, relief pressure, and ow capacity are determined,
we can identify a number of potentially suitable relief valves
using the Quick Selection Guide in the front of each application
section in this application guide. These selection guides give
relief set (inlet) pressures, capacities, and type numbers. These
guides can then be further narrowed by reviewing individual
product pages in each section.
Final Selection
Final selection is usually a matter of compromise. Relief
capacities, build-up levels, sensitivity, throttling capabilities,
cost of installation and maintenance, space requirements, initial
purchase price, and other attributes are all considered when
choosing any relief valve.
Applicable Regulations
The relief valves installed in some applications must meet
governmental, industry, or company criteria.
Sizing and Selection Exercise
To gain a better understanding of the selection and sizing
process, it may be helpful to step through a typical relief valve
sizing exercise.
Initial Parameters
We’ll assume that we need to specify an appropriate relief valve
for a regulator serving a large plant air supply. There is sufcient
space to install the relief valve and the controlled uid is clean
plant air that can be exhausted without adding a vent stack.
Performance Considerations
The plant supervisor wants the relief valve to throttle open
smoothly so that pressure surges will not damage instruments
and equipment in the downstream system. This will require the
selection of a relief valve that will open smoothly. Plant equipment
is periodically shut down but the air supply system operates
continuously. Therefore, the relief valve must also have the
capacity to exhaust the full ow of the upstream system.
Upstream Regulator
The regulator used is 1-inch in size with a 3/8-inch orice. The
initial system parameters of pressure and ow were determined
when the regulator was sized for this application.
Pressure Limits
The plant maintenance engineer has determined that the relief
valve should begin to open at 20 psig, and downstream pressure
should not rise above 30 psig maximum allowable system pressure.
Relief Valve Flow Capacity
The wide-open regulator ow is calculated to be 23 188 SCFH.
Page 46
Principles of Relief Valves
616
Te c h n i c a l
Relief Valve Selection
Quick Selection Guide
Find the Relief Valve Quick Selection Guide in this Application
Guide; it gives relief set (inlet) pressures and comparative ow
capacities of various relief valves. Because this guide is used to
identify potentially suitable relief valves, we can check the relief
set (inlet) pressures closest to 20 psig and narrow the range of
choices. We nd that two relief valves have the required ow
capacity at our desired relief set (inlet) pressure.
Product Pages
If we look at the product pages for the potential relief valves, we
nd that a 1-inch Type 289H provides the required capacity within
the limits of pressure build-up specied in our initial parameters.
Checking Capacity
Capacity curves for the 1-inch Type 289H with this spring are
shown in Figure 10. By following the curve for the 20 psig
setpoint to the point where it intersects with the 30 psig
division, we nd that our relief valve can handle more than the
23 188 SCFH required.
CAPACITIES IN THOUSANDS OF SCFH (Nm³/h) OF AIR
Figure 10. Type 289H Flow Capacities
INLET PRESSURE, PSIG
INLET PRESSURE, bar
0
0
7.75
(0,208)
15.5
(0,415)
23.3
(0,624)
31
(0,831)
38.8
(1,04)
46.5
(1,25)
54.3
(1,46)
62
(1,66)
40
50
70
60
10
30
20
0
2,8
3,4
4,8
4,1
0,69
2,1
1,4
1-INCH TYPE 289H
VENT SCREEN INSTALLED
15 PSIG (1,0 bar)
1D751527022
10 PSIG (0,69 bar)
1D751527022
1 PSIG (0,069 bar)
1F782697052
4 PSIG (0,28 bar)
40 PSIG (2,8 bar)
1D7455T0012
50 PSIG (3,4 bar)
1D7455T0012
30 PSIG (2,1 bar)
1D7455T0012
20 PSIG (1,4 bar)
1D751527022
B2309
Page 47
Principles of Series Regulation and Monitor Regulators
617
Te c h n i c a l
Series Regulation
Series regulation is one of the simplest systems used to provide
overpressure protection by containment. In the example shown in
Figure 1, the inlet pressure is 100 psig, the desired downstream pressure
is 10 psig, and the maximum allowable operating pressure (MAOP) is
40 psig. The setpoint of the downstream regulator is 10 psig, and the
setpoint of the upstream regulator is 30 psig.
Because of the problem in maintaining close control of P2, series
regulation is best suited to applications where the regulator station
is reducing pressure to a value substantially below the maximum
allowable operating pressure of the downstream system. Farm taps
are a good example. The problem of low-pressure drop across the
second regulator is less pronounced in low ow systems.
Upstream Wide-Open Monitors
The only difference in conguration between series regulation
and monitors is that in monitor installations, both regulators sense
downstream pressure, P2. Thus, the upstream regulator must have
a control line.
Figure 1. Series Regulation
Failed System Response
If regulator B fails, downstream pressure (P2) is maintained at the
setpoint of regulator A less whatever drop is required to pass the
required ow through the failed regulator B. If regulator A fails,
the intermediate pressure will be 100 psig. Regulator B must be
able to withstand 100 psig inlet pressure.
Regulator Considerations
Either direct-operated or pilot-operated regulators may be used
in this system. Should regulator A fail, P
Intermediate
will approach P1
so the outlet rating and spring casing rating of regulator A must
be high enough to withstand full P1. This situation may suggest
the use of a relief valve between the two regulators to limit the
maximum value of P
Intermediate
.
Applications and Limitations
A problem with series regulation is maintaining tight control of P2
if the downstream regulator fails. In this arrangement, it is often
impractical to have the setpoints very close together. If they are,
the pressure drop across regulator B will be quite small. With a
small pressure drop, a very large regulator may be required to pass
the desired ow.
System Values
In the example shown in Figure 2, assume that P1 is 100 psig, and
the desired downstream pressure, P2, is 10 psig. Also assume that
the maximum allowable operating pressure of the downstream
system is 20 psig; this is the limit we cannot exceed. The setpoint
of the downstream regulator is set at 10 psig to maintain the desired
P2 and the setpoint of the upstream regulator is set at 15 psig to
maintain P2 below the maximum allowable operating pressure.
Normal Operation
When both regulators are functioning properly, regulator B holds
P2 at its setpoint of 10 psig. Regulator A, sensing a pressure lower
than its setpoint of 15 psig tries to increase P2 by going wide-open.
This conguration is known as an upstream wide-open monitor
where upstream regulator A monitors the pressure established by
regulator B. Regulator A is referred to as the monitor or standby
regulator while regulator B is called the worker or the operator.
SETPOINT = 30 PSIG
SETPOINT = 10 PSIG
P
1
100 PSIG
P
2
10 PSIG
A B
P
INTERMEDIATE
30 PSIG
SETPOINT = 10 PSIG
SETPOINT = 15 PSIG
P
1
100 PSIG
P
2
10 PSIG
P
INTERMEDIATE
A
B
MONITOR REGULATOR WORKER REGULATOR
IN WIDE-OPEN MONITOR SYSTEMS, BOTH REGULATORS SENSE DOWNSTREAM PRESSURE. SETPOINTS
MAY BE VERY CLOSE TO EACH OTHER. IF THE WORKER REGULATOR FAILS, THE MONITOR ASSUMES
CONTROL AT A SLIGHTLY HIGHER SETPOINT. IF THE MONITOR REGULATOR FAILS, THE WORKER
CONTINUES TO PROVIDE CONTROL.
Figure 2. Wide-Open Upstream Monitor
Page 48
Principles of Series Regulation and Monitor Regulators
618
Te c h n i c a l
Worker Regulator B Fails
If regulator B fails open, regulator A, the monitor, assumes control
and holds P2 at 15 psig. Note that pressure P
Intermediate
is now P2 plus
whatever drop is necessary to pass the required ow through the
failed regulator B.
Equipment Considerations
Wide-open monitoring systems may use either direct- or pilotoperated regulators, the choice of which is dependent on other
system requirements. Obviously, the upstream regulator must
have external registration capability in order to sense downstream
pressure, P2.
In terms of ratings, P
Intermediate
will rise to full P1 when regulator A
fails, so the body outlet of regulator A and the inlet of regulator B
must be rated for full P1.
Downstream Wide-Open Monitors
The difference between upstream and downstream monitor systems
(Figure 3) is that the functions of the two regulators are reversed.
In other words, the monitor, or standby regulator, is downstream
of the worker, or operator. Systems can be changed from upstream
to downstream monitors, and vice-versa, by simply reversing the
setpoints of the two regulators.
Worker Regulator A Fails
If the worker, regulator A, fails in an open position, the monitor,
regulator B, senses the increase in P2 and holds P2 at its setpoint
of 15 psig. Note that P
Intermediate
is now P1 minus whatever drop is
taken across the failed regulator A.
Upstream Versus Downstream Monitors
The decision to use either an upstream or downstream monitor
system is largely a matter of personal preference or company policy.
In normal operation, the monitor remains open while the worker is
frequently exercised. Many users see value in changing the system
from an upstream to a downstream monitor at regular intervals,
much like rotating the tires on an automobile. Most uids have
some impurities such as moisture, rust, or other debris, which may
deposit on regulator components, such as stems, and cause them to
become sticky or bind. Therefore, occasionally reversing the roles
of the regulators so that both are exercised is sometimes seen as a
means of ensuring that protection is available when needed. The
job of switching is relatively simple as only the setpoints of the two
regulators are changed. In addition, the act of changing from an
upstream to a downstream monitor requires that someone visit the
site so there is an opportunity for routine inspection.
Working Monitors
Working monitors (Figure 4) use design elements from both
series regulation and wide-open monitors. In a working monitor
installation, the two regulators are continuously working as series
regulators to take two pressure cuts.
THE ONLY DIFFERENCE BETWEEN UPSTREAM WIDE-OPEN MONITOR SYSTEMS AND DOWNSTREAM
WIDE-OPEN MONITOR SYSTEMS IS THE ROLE EACH REGULATOR PLAYS. WORKERS AND MONITORS
MAY BE SWITCHED BY SIMPLY REVERSING THE SETPOINTS.
Figure 3. Wide-Open Downstream Monitor
Normal Operation
Again, assume an inlet pressure of 100 psig and a controlled
pressure (P2) of 10 psig. Regulator A is now the worker so it
maintains P2 at its setpoint of 10 psig. Regulator B, the monitor, is
set at 15 psig and so remains open.
SETPOINT = 10 PSIG
MONITOR PILOT
(SETPOINT —15 PSIG)
P
1
100 PSIG
P
2
10 PSIG
P
INTERMEDIATE
45 PSIG
WORKER PILOT
(SETPOINT—45 PSIG)
Figure 4. Working Monitor
WORKING MONITOR SYSTEMS MUST USE A PILOT-OPERATED REGULATOR AS THE MONITOR, WHICH IS
ALWAYS IN THE UPSTREAM POSITION. TWO PILOTS ARE USED ON THE MONITOR REGULATOR; ONE TO
CONTROL THE INTERMEDIATE PRESSURE AND ONE TO MONITOR THE DOWNSTREAM PRESSURE. BY
TAKING TWO PRESSURE DROPS, BOTH REGULATORS ARE ALLOWED TO EXERCISE.
SETPOINT = 15 PSIG
SETPOINT = 10 PSIG
P
1
100 PSIG
P
2
10 PSIG
P
INTERMEDIATE
A
B
MONITOR REGULATORWORKER REGULATOR
Page 49
Principles of Series Regulation and Monitor Regulators
619
Te c h n i c a l
Downstream Regulator
The downstream regulator may be either direct or pilot-operated.
It is installed just as in a series or wide-open monitor system. Its
setpoint controls downstream pressure, P2.
Upstream Regulator
The upstream regulator must be a pilot-operated type because it
uses two pilots; a monitor pilot and a worker pilot. The worker
pilot is connected just as in series regulation and controls the
intermediate pressure P
Intermediate
. Its setpoint (45 psig) is at some
intermediate value that allows the system to take two pressure
drops. The monitor pilot is in series ahead of the worker pilot
and is connected so that it senses downstream pressure, P2. The
monitor pilot setpoint (15 psig) is set slightly higher than the
normal P2 (10 psig).
Normal Operation
When both regulators are performing properly, downstream
pressure is below the setting of the monitor pilot, so it is fully open
trying to raise system pressure. Standing wide-open, the monitor
pilot allows the worker pilot to control the intermediate pressure,
P
Intermediate
at 45 psig. The downstream regulator is controlling
P2 at 10 psig.
Downstream Regulator Fails
If the downstream regulator fails, the monitor pilot will sense the
increase in pressure and take control at 15 psig.
Upstream Regulator Fails
If the upstream regulator fails, the downstream regulator will remain
in control at 10 psig. Note that the downstream regulator must
be rated for the full system inlet pressure P1 of 100 psig because
this will be its inlet pressure if the upstream regulator fails. Also
note that the outlet rating of the upstream regulator, and any other
components that are exposed to P
Intermediate
, must be rated for full P1.
Sizing Monitor Regulators
The difcult part of sizing monitor regulators is that P
Intermediate
is needed to determine the ow capacity for both regulators.
Because P
Intermediate
is not available, other sizing methods are used
to determine the capacity. There are three methods for sizing
monitor regulators: estimating ow when pressure drop is critical,
assuming P
Intermediate
to calculate ow, and the Fisher® Monitor
Sizing Program.
Estimating Flow when Pressure Drop is Critical
If the pressure drop across both regulators from P1 to P2 is critical
(assume P
Intermediate
= P1 - P2/2 + P2, P1 - P
Intermediate
> P1, and P
Intermediate
- P2 > 1/2 P
Intermediate
), and both regulators are the same type, the
capacity of the two regulators together is 70 to 73% of a single
regulator reducing the pressure from P1 to P2.
Assuming P
Intermediate
to Determine Flow
Assume P
Intermediate
is halfway between P1 and P2. Guess a regulator
size. Use the assumed P
Intermediate
and the Cg for each regulator to
calculate the available ow rate for each regulator. If P
Intermediate
was correct, the calculated ow through each regulator will be the
same. If the ows are not the same, change P
Intermediate
and repeat
the calculations. (P
Intermediate
will go to the correct assumed pressure
whenever the ow demand reaches maximum capacity.)
Fisher® Monitor Sizing Program
Emerson Process Management - Regulator Technologies offers a
Monitor Sizing Program on the Regulator Technologies Literature
CD. Call your local Sales Ofce to request a copy of the CD. To
locate your local Sales Ofce, log on to: www.emersonprocess.
com/regulators.
Page 50
Vacuum Control
620
Te c h n i c a l
Vacuum Applications
Vacuum regulators and vacuum breakers are widely used in process
plants. Conventional regulators and relief valves might be suitable
for vacuum service if applied correctly. This section provides
fundamentals and examples.
Vacuum Control Devices
Just like there are pressure reducing regulators and pressure relief
valves for positive pressure service, there are also two basic
types of valves for vacuum service. The terms used for each are
sometimes confusing. Therefore, it is sometimes necessary to ask
further questions to determine the required function of the valve.
The terms vacuum regulator and vacuum breaker will be used in
these pages to differentiate between the two types.
Vacuum Regulators
Vacuum regulators maintain a constant vacuum at the regulator
inlet. A loss of this vacuum (increase in absolute pressure) beyond
setpoint registers on the diaphragm and opens the disk. It depends
on the valve as to which side of the diaphragm control pressure is
measured. Opening the valve plug permits a downstream vacuum
of lower absolute pressure than the controlled vacuum to restore
the upstream vacuum to its original setting.
Besides the typical vacuum regulator, a conventional regulator can
be suitable if applied correctly. Any pressure reducing regulator
(spring to open device) that has an external control line connection
and an O-ring stem seal can be used as a vacuum regulator.
Installation requires a control line to connect the vacuum being
controlled and the spring case. The regulator spring range is now a
negative pressure range and the body ow direction is the same as
in conventional pressure reducing service.
Vacuum Breakers (Relief Valves)
Vacuum breakers are used in applications where an increase in
vacuum must be limited. An increase in vacuum (decrease in
absolute pressure) beyond a certain value causes the diaphragm to
move and open the disk. This permits atmospheric pressure or a
positive pressure, or an upstream vacuum that has higher absolute
pressure than the downstream vacuum, to enter the system and
restore the controlled vacuum to its original pressure setting.
A vacuum breaker is a spring-to-close device, meaning that if there
is no pressure on the valve the spring will push the valve plug into
its seat. There are various Fisher® brand products to handle this
application. Some valves are designed as vacuum breakers.
Fisher brand relief valves can also be used as vacuum breakers.
ABSOLUTE
ZERO
-14.7 PSIG (-1,01 bar g),
0 PSIA (0 bar a)
ATMOSPHERIC
5 PSIG (0,34 bar g) VACUUM,
-5 PSIG (-0,34 bar g),
9.7 PSIA (0,67 bar a)
0 PSIG (0 bar g),
14.7 PSIA (1,01 bar a)
5 PSIG (0,34 bar g),
19.7 PSIA (1,36 bar a)
POSITIVE PRESSURE
VACUUM
1 PSIG (0,069 bar) = 27.7-INCHES OF WATER (69 mbar) = 2.036-INCHES OF MERCURY
1kg/cm2 = 10.01 METERS OF WATER = 0.7355 METERS OF MERCURY
Figure 1. Vacuum Terminology
Vacuum Terminology
Engineers use a variety of terms to describe vacuum, which
can cause some confusion. Determine whether the units are in
absolute pressure or gauge pressure (0 psi gauge (0 bar gauge) is
atmospheric pressure).
• 5 psig (0,34 bar g) vacuum is 5 psi (0,34 bar) below
atmospheric pressure.
• -5 psig (-0,34 bar g) is 5 psi (0,34 bar) below
atmospheric pressure.
• 9.7 psia (0,67 bar a) is 9.7 psi (0,67 bar) above absolute zero
or 5 psi (0,34 bar) below atmospheric pressure
(14.7 psia - 5 psi = 9.7 psia (1,01 bar a - 0,34 bar = 0,67 bar a)).
Page 51
Vacuum Control
621
Te c h n i c a l
A conventional relief valve can be used as a vacuum breaker, as
long as it has a threaded spring case vent so a control line can be
attached. If inlet pressure is atmospheric air, then the internal
pressure registration from body inlet to lower casing admits
atmospheric pressure to the lower casing. If inlet pressure is not
atmospheric, a relief valve in which the lower casing can be vented
to atmosphere when the body inlet is pressurized must be chosen.
In this case, the terminology “blocked throat” and “external
registration with O-ring stem seal” are used for clarity.
UNAVOIDABLE
LEAKAGE
TYPE Y695VR VACUUM REGULATOR
POSITIVE PRESSURE OR
ATMOSPHERE, OR A LESSER VACUUM
THAN THE VACUUM BEING LIMITED
TYPE Y690VB VACUUM BREAKER
Figure 3. Typical Vacuum Breaker
Figure 2. Typical Vacuum Regulator
INLET PRESSURE
CONTROL PRESSURE (VACUUM)
ATMOSPHERIC PRESSURE
A spring that normally has a range of 6 to 11-inches w.c. (15 to 27 mbar)
positive pressure will now have a range of 6 to 11-inches w.c.
(15 to 27 mbar) vacuum (negative pressure). It may be expedient to
bench set the vacuum breaker if the type chosen uses a spring case
closing cap. Removing the closing cap to gain access to the adjusting
screw will admit air into the spring case when in vacuum service.
CONTROL PRESSURE (VACUUM)
OUTLET PRESSURE (VACUUM)
ATMOSPHERIC PRESSURE
VACUUM
BEING CONTROLLED
VACUUM
PUMP
HIGHER
VACUUM SOURCE
VACUUM
PUMP
VACUUM
BEING LIMITED
B2582
B2583
Page 52
Vacuum Control
622
Te c h n i c a l
VACUUM BEING CONTROLLED HIGHER VACUUM SOURCE
ATMOSPHERIC
BLEED
VACUUM
PUMP
FIXED
RESTRICTION
TYPE Y695VRM
VACUUM REGULATOR
TYPE 1098-EGR
PRESSURE REDUCING REGULATOR
Figure 5. Type Y695VRM used with Type 1098-EGR in a Vacuum Regulator Installation
Figure 4. Type 133L
VACUUM BEING
REGULATED
VACUUM SOURCE
Vacuum Regulator Installation Examples
CONTROL PRESSURE (VACUUM)
ATMOSPHERIC PRESSURE
OUTLET PRESSURE (VACUUM)
CONTROL PRESSURE (VACUUM)
LOADING PRESSURE
ATMOSPHERIC PRESSURE
OUTLET PRESSURE (VACUUM)
A6555
Page 53
Vacuum Control
623
Te c h n i c a l
CONSTANT INLET OR
ATMOSPHERIC PRESSURE
TYPE 1098-EGR
PRESSURE REDUCING REGULATOR
FIXED
RESTRICTION
TYPE Y690VB
VACUUM BREAKER
VACUUM TANK
VACUUM
PUMP
Figure 7. Type Y690VB used with Type 1098-EGR in a Vacuum Breaker Installation. If the positive pressure exceeds the
Type 1098-EGR casing rating, then a Type 67CF with a Type H800 relief valve should be added.
VACUUM TANK
VACUUM
PUMP
ATMOSPHERIC OR
CONSTANT INLET ONLY
Figure 6. Type 1805
Vacuum Breaker Installation Examples
INLET PRESSURE
CONTROL PRESSURE (VACUUM)
INLET PRESSURE
LOADING PRESSURE
ATMOSPHERIC PRESSURE
CONTROL PRESSURE (VACUUM)
VACUUM BEING LIMITED
A6671
Page 54
Vacuum Control
624
Te c h n i c a l
M1047
Type 289H
VACUUM TANK
ATMOSPHERIC
INLET ONLY
PITOT TUBE SERVES TO REGISTER
VACUUM ON TOP OF DIAPHRAGM
(NO CONTROL LINE NEEDED)
TYPE 66R
VACUUM
TANK
VACUUM
PUMP
VACUUM
PUMP
ATMOSPHERIC INLET
PRESSURE
Figure 8. Type 289H Relief Valve used in a Vacuum Breaker Installation
If inlet is positive pressure:
• Select balancing diaphragm and tapped lower
casing construction.
• Leave lower casing open to atmospheric pressure.
Figure 10. Type 66RR Relief Valve used in a Vacuum Breaker Installation
VACUUM
TANK
MAIN VALVE
PLUG
TYPE 66RR
LOADING PRESSURE
CONTROLLED PRESSURE (VACUUM)
INLET PRESSURE
MAIN VALVE
DIAPHRAGM
PILOT VALVE DISK
FIXED RESTRICTION
PILOT CONTROL
SPRING
Vacuum Breaker Installation Examples
Figure 9. Type 66R Relief Valve used in a Vacuum Breaker Installation
INLET PRESSURE
CONTROL PRESSURE (VACUUM)
INLET PRESSURE
CONTROL PRESSURE (VACUUM)
M1047
M1041
M1056
Page 55
Vacuum Control
625
Te c h n i c a l
Gas Blanketing in Vacuum
When applications arise where the gas blanketing requirements
are in vacuum, a combination of a vacuum breaker and a regulator
may be used. For example, in low inches of water column vacuum,
a Type Y690VB vacuum breaker and a Type 66-112 vacuum
regulator can be used for very precise control.
Vacuum blanketing is useful for vessel leakage to atmosphere
and the material inside the vessel is harmful to the surrounding
environment. If leakage were to occur, only outside air would
enter the vessel because of the pressure differential in the tank.
Therefore, any process vapors in the tank would be contained.
Features of Fisher® Brand Vacuum Regulators
and Breakers
• Precision Control of Low Pressure Settings—Large
diaphragm areas provide more accurate control at low
pressure settings. Some of these regulators are used as pilots
on our Tank Blanketing and Vapor Recovery Regulators.
Therefore, they are designed to be highly accurate, usually
within 1-inch w.c. (2 mbar).
Figure 11. Example of Gas Blanketing in Vacuum
20 TO 30 PSIG (1,4 to 2,1 bar)
NITROGEN OR OTHER GAS
VACUUM BREAKER TYPE Y690VB
SET TO OPEN AT
-2-INCHES W.C. (-5 mbar)
TO EXHAUST HEADER
-9-INCHES W.C. (-22 mbar)
VACUUM REGULATOR
TYPE 66-112 SET TO CONTROL
AT -1-INCH W.C. (-2 mbar) VACUUM
CONTROL LINES
TANK — 1-INCH W.C. (2 mbar) VACUUM
• Corrosion Resistance—Constructions are available in a
variety of materials for compatibility with corrosive process
gases. Wide selection of elastomers compatible with
owing media.
• Rugged Construction—Diaphragm case and internal parts are
designed to withstand vibration and shock.
• Wide Product Offering—Fisher
®
brand regulators can be
either direct-operated or pilot-operated regulators.
• Fisher Brand Advantage—Widest range of products and
a proven history in the design and manufacture of process
control equipment. A sales channel that offers local stock
and support.
• Spare Parts—Low cost parts that are interchangeable with
other Fisher brand in your plant.
• Easy Sizing and Selection—Most applications can be
sized utilizing the Fisher brand Sizing Program and
Sizing Coefcients.
BJ9004
Page 56
Valve Sizing Calculations (Traditional Method)
626
Te c h n i c a l
Introduction
Fisher® regulators and valves have traditionally been sized using
equations derived by the company. There are now standardized
calculations that are becoming accepted worldwide. Some product
literature continues to demonstrate the traditional method, but
the trend is to adopt the standardized method. Therefore, both
methods are covered in this application guide.
Improper valve sizing can be both expensive and inconvenient.
A valve that is too small will not pass the required ow, and
the process will be starved. An oversized valve will be more
expensive, and it may lead to instability and other problems.
The days of selecting a valve based upon the size of the pipeline
are gone. Selecting the correct valve size for a given application
requires a knowledge of process conditions that the valve will
actually see in service. The technique for using this information
to size the valve is based upon a combination of theory
and experimentation.
Sizing for Liquid Service
Using the principle of conservation of energy, Daniel Bernoulli
found that as a liquid ows through an orice, the square of the
uid velocity is directly proportional to the pressure differential
across the orice and inversely proportional to the specic gravity
of the uid. The greater the pressure differential, the higher the
velocity; the greater the density, the lower the velocity. The
volume ow rate for liquids can be calculated by multiplying the
uid velocity times the ow area.
By taking into account units of measurement, the proportionality
relationship previously mentioned, energy losses due to friction
and turbulence, and varying discharge coefcients for various
types of orices (or valve bodies), a basic liquid sizing equation
can be written as follows
Q = CV ∆P / G
(1)
where:
Q = Capacity in gallons per minute
Cv = Valve sizing coefcient determined experimentally for
each style and size of valve, using water at standard
conditions as the test uid
∆P = Pressure differential in psi
G = Specic gravity of uid (water at 60°F = 1.0000)
Thus, Cv is numerically equal to the number of U.S. gallons of
water at 60°F that will ow through the valve in one minute when
the pressure differential across the valve is one pound per square
inch. Cv varies with both size and style of valve, but provides an
index for comparing liquid capacities of different valves under a
standard set of conditions.
To aid in establishing uniform measurement of liquid ow capacity
coefcients (Cv) among valve manufacturers, the Fluid Controls
Institute (FCI) developed a standard test piping arrangement,
shown in Figure 1. Using such a piping arrangement, most
valve manufacturers develop and publish Cv information for
their products, making it relatively easy to compare capacities of
competitive products.
To calculate the expected Cv for a valve controlling water or other
liquids that behave like water, the basic liquid sizing equation
above can be re-written as follows
CV = Q
G
∆P
(2)
Viscosity Corrections
Viscous conditions can result in signicant sizing errors in using
the basic liquid sizing equation, since published Cv values are
based on test data using water as the ow medium. Although the
majority of valve applications will involve uids where viscosity
corrections can be ignored, or where the corrections are relatively
small, uid viscosity should be considered in each valve selection.
Emerson Process Management has developed a nomograph
(Figure 2) that provides a viscosity correction factor (Fv). It can
be applied to the standard Cv coefcient to determine a corrected
coefcient (Cvr) for viscous applications.
Finding Valve Size
Using the Cv determined by the basic liquid sizing equation and
the ow and viscosity conditions, a uid Reynolds number can be
found by using the nomograph in Figure 2. The graph of Reynolds
number vs. viscosity correction factor (Fv) is used to determine
the correction factor needed. (If the Reynolds number is greater
than 3500, the correction will be ten percent or less.) The actual
required Cv (Cvr) is found by the equation:
C
vr
= FV CV (3)
From the valve manufacturer’s published liquid capacity
information, select a valve having a Cv equal to or higher than the
required coefcient (Cvr) found by the equation above.
Figure 1. Standard FCI Test Piping for Cv Measurement
PRESSURE
INDICATORS
∆P ORIFICE
METER
INLET VALVE TEST VALVE LOAD VALVE
FLOW
Page 57
627
Te c h n i c a l
Valve Sizing Calculations (Traditional Method)
3000
3,000
4,000
6,000
8,000
10,000
20,000
30,000
40,000
60,000
80,000
100,000
200,000
300,000
400,000
600,000
800,000
C
INDEX
1 2 3 4 6 8 10 20 30 40 60 80 100 200
2000
2,000
1,000
2,000
2,000
3,000
3,000
4,000
4,000
6,000
6,000
8,000
8,000
10,000
10,000
800
600
400
300
200
100
80
60
40
30
20
10
8
6
4
3
2
1
1,000
1,000
2,000
2,000
3,000
3,000
4,000
4,000
6,000
6,000
8,000
8,000
10,000
10,000
20,000
20,000
30,000
30,000
40,000
40,000
60,000
60,000
80,000
80,000
100,000
100,000
200,000
300,000
400,000
800
800
600
600
400
400
300
300
200
200
100
100
80
80
60
60
40
40
35
32.6
30
20
10
8
6
4
3
2
1
0.8
0.6
0.4
0.3
0.2
0.1
0.08
0.06
0.04
0.03
0.02
0.01
1,000
1,000
800
600
800
600
400
300
200
100
80
60
40
30
20
400
300
200
100
80
60
40
30
20
10
10
8
8
6
6
4
4
3
3
2
2
1
0.8
0.6
0.4
0.2
0.1
0.3
1
0.8
0.6
0.4
0.3
0.2
0.1
0.08
0.06
0.04
0.04
0.06
0.08
0.03
0.03
0.02
0.02
0.01
0.01
0.008
0.008
0.006
0.006
0.004
0.004
0.003
0.003
0.002
0.002
0 .001
0.001
0.0008
0.0008
0.0006
0.0006
0.0004
0.0004
0.0003
0.0003
0.0002
0.0002
0.0001
0.0001
1000
800
600
400
300
200
100
80
60
40
30
20
10
8
6
4
3
2
1
0.8
0.6
0.4
0.3
0.2
0.1
0.08
0.06
0.04
0.03
0.02
0.01
Figure 2. Nomograph for Determining Viscosity Correction
Nomograph Instructions
Use this nomograph to correct for the effects of viscosity. When
assembling data, all units must correspond to those shown on the
nomograph. For high-recovery, ball-type valves, use the liquid
ow rate Q scale designated for single-ported valves. For buttery
and eccentric disk rotary valves, use the liquid ow rate Q scale
designated for double-ported valves.
Nomograph Equations
1. Single-Ported Valves:
NR = 17250
Q
C
V νCS
2. Double-Ported Valves:
NR = 12200
Q
C
V νCS
Nomograph Procedure
1. Lay a straight edge on the liquid sizing coefcient on C
v
scale and ow rate on Q scale. Mark intersection on index
line. Procedure A uses value of Cvc; Procedures B and C use
value of Cvr.
2. Pivot the straight edge from this point of intersection with
index line to liquid viscosity on proper n scale. Read Reynolds
number on NR scale.
3. Proceed horizontally from intersection on NR scale to proper
curve, and then vertically upward or downward to Fv scale.
Read Cv correction factor on Fv scale.
3,000
4,000
6,000
8,000
10,000
20,000
30,000
40,000
60,000
80,000
100,000
200,000
300,000
400,000
600,000
800,000
1,000,000
1 2 3 4 6
8 10 20 30 40 60 80 100 200
1 2 3 4 6 8 10 20 30 40 60 80 100 200
2,000
1,000
1,000
2,000
2,000
3,000
3,000
4,000
4,000
6,000
6,000
8,000
8,000
10,000
10,000
20,000
20,000
30,000
30,000
40,000
40,000
60,000
60,000
80,000
80,000
100,000
100,000
200,000
300,000
400,000
800
800
600
600
400
400
300
300
200
200
100
100
80
80
60
60
40
40
35
32.6
30
20
10
8
6
4
3
2
1
1,000
800
600
400
300
200
100
80
60
40
30
20
10
8
6
4
3
2
1
0.8
0.6
0.4
0.2
0.1
0.3
0.04
0.06
0.08
0.03
0.02
0.01
LIQUID FLOW COEFFICIENT, C
V
C
V
LIQUID FLOW RATE (SINGLE PORTED ONLY), GPM
Q
LIQUID FLOW RATE (DOUBLE PORTED ONLY), GPM
KINEMATIC VISCOSITY V
CS
- CENTISTOKES
VISCOSITY - SAYBOLT SECONDS UNIVERSAL
REYNOLDS NUMBER - N
R
C
V
CORRECTION FACTOR, F
V
INDEX
C
V
CORRECTION FACTOR, F
V
H
R
F
V
FOR PREDICTING PRESSURE DROP
FOR SELECTING VALVE SIZE
FOR PREDICTING
FLOW RATE
3,000
4,000
6,000
8,000
10,000
20,000
30,000
40,000
60,000
80,000
100,000
200,000
INDEX
1 2 3 4 6 8 10 20 30 40 60 80 100 200
2,000
1,000
1,000
2,000
2,000
3,000
3,000
4,000
4,000
6,000
6,000
8,000
8,000
10,000
10,000
20,000
20,000
30,000
30,000
40,000
40,000
60,000
60,000
80,000
80,000
100,000
100,000
200,000
300,000
400,000
800
800
600
600
400
400
300
300
200
200
100
100
80
80
60
60
40
40
35
32.6
30
20
10
8
6
4
3
2
1
1,000
800
600
400
300
200
100
80
60
40
30
20
10
8
6
4
3
2
1
0.8
0.6
0.4
0.2
0.1
0.3
0.04
0.06
0.08
0.03
0.02
0.01
Page 58
Valve Sizing Calculations (Traditional Method)
628
Te c h n i c a l
Predicting Flow Rate
Select the required liquid sizing coefcient (Cvr) from the
manufacturer’s published liquid sizing coefcients (Cv) for the
style and size valve being considered. Calculate the maximum
ow rate (Q
max
) in gallons per minute (assuming no viscosity
correction required) using the following adaptation of the basic
liquid sizing equation:
Q
max
= Cvr ΔP / G (4)
Then incorporate viscosity correction by determining the uid
Reynolds number and correction factor Fv from the viscosity
correction nomograph and the procedure included on it.
Calculate the predicted ow rate (Q
pred
) using the formula:
Q
pred
=
Q
max
F
V
(5)
Predicting Pressure Drop
Select the required liquid sizing coefcient (Cvr) from the published
liquid sizing coefcients (Cv) for the valve style and size being
considered. Determine the Reynolds number and correct factor Fv
from the nomograph and the procedure on it. Calculate the sizing
coefcient (Cvc) using the formula:
CVC =
C
vr
F
v
(6)
Calculate the predicted pressure drop (∆P
pred
) using the formula:
ΔP
pred
= G (Q/Cvc)2 (7)
Flashing and Cavitation
The occurrence of ashing or cavitation within a valve can have a
signicant effect on the valve sizing procedure. These two related
physical phenomena can limit ow through the valve in many
applications and must be taken into account in order to accurately
size a valve. Structural damage to the valve and adjacent piping
may also result. Knowledge of what is actually happening within
the valve might permit selection of a size or style of valve which
can reduce, or compensate for, the undesirable effects of ashing
or cavitation.
The “physical phenomena” label is used to describe ashing and
cavitation because these conditions represent actual changes in
the form of the uid media. The change is from the liquid state
to the vapor state and results from the increase in uid velocity at
or just downstream of the greatest ow restriction, normally the
valve port. As liquid ow passes through the restriction, there is a
necking down, or contraction, of the ow stream. The minimum
cross-sectional area of the ow stream occurs just downstream of
the actual physical restriction at a point called the vena contracta,
as shown in Figure 3.
To maintain a steady ow of liquid through the valve, the velocity
must be greatest at the vena contracta, where cross sectional
area is the least. The increase in velocity (or kinetic energy) is
accompanied by a substantial decrease in pressure (or potential
energy) at the vena contracta. Farther downstream, as the uid
stream expands into a larger area, velocity decreases and pressure
increases. But, of course, downstream pressure never recovers
completely to equal the pressure that existed upstream of the
valve. The pressure differential (∆P) that exists across the valve
Figure 3. Vena Contracta
Figure 4. Comparison of Pressure Profiles for
High and Low Recovery Valves
VENA CONTRACTA
RESTRICTION
FLOW
FLOW
P
1
P
2
P
1
P
2
P
2
P
2
HIGH RECOVERY
LOW RECOVERY
P
1
Page 59
629
Te c h n i c a l
Valve Sizing Calculations (Traditional Method)
is a measure of the amount of energy that was dissipated in the
valve. Figure 4 provides a pressure prole explaining the differing
performance of a streamlined high recovery valve, such as a ball
valve and a valve with lower recovery capabilities due to greater
internal turbulence and dissipation of energy.
Regardless of the recovery characteristics of the valve, the pressure
differential of interest pertaining to ashing and cavitation is the
differential between the valve inlet and the vena contracta. If
pressure at the vena contracta should drop below the vapor pressure
of the uid (due to increased uid velocity at this point) bubbles
will form in the ow stream. Formation of bubbles will increase
greatly as vena contracta pressure drops further below the vapor
pressure of the liquid. At this stage, there is no difference between
ashing and cavitation, but the potential for structural damage to
the valve denitely exists.
If pressure at the valve outlet remains below the vapor pressure
of the liquid, the bubbles will remain in the downstream system
and the process is said to have “ashed.” Flashing can produce
serious erosion damage to the valve trim parts and is characterized
by a smooth, polished appearance of the eroded surface. Flashing
damage is normally greatest at the point of highest velocity, which
is usually at or near the seat line of the valve plug and seat ring.
However, if downstream pressure recovery is sufcient to raise the
outlet pressure above the vapor pressure of the liquid, the bubbles
will collapse, or implode, producing cavitation. Collapsing of the
vapor bubbles releases energy and produces a noise similar to what
one would expect if gravel were owing through the valve. If the
bubbles collapse in close proximity to solid surfaces, the energy
released gradually wears the material leaving a rough, cylinder
like surface. Cavitation damage might extend to the downstream
pipeline, if that is where pressure recovery occurs and the bubbles
collapse. Obviously, “high recovery” valves tend to be more
subject to cavitation, since the downstream pressure is more likely
to rise above the vapor pressure of the liquid.
Choked Flow
Aside from the possibility of physical equipment damage due to
ashing or cavitation, formation of vapor bubbles in the liquid ow
stream causes a crowding condition at the vena contracta which
tends to limit ow through the valve. So, while the basic liquid
sizing equation implies that there is no limit to the amount of ow
through a valve as long as the differential pressure across the valve
increases, the realities of ashing and cavitation prove otherwise.
If valve pressure drop is increased slightly beyond the point where
bubbles begin to form, a choked ow condition is reached. With
constant upstream pressure, further increases in pressure drop (by
reducing downstream pressure) will not produce increased ow.
The limiting pressure differential is designated ∆P
allow
and the valve
recovery coefcient (Km) is experimentally determined for each
valve, in order to relate choked ow for that particular valve to the
basic liquid sizing equation. Km is normally published with other
valve capacity coefcients. Figures 5 and 6 show these ow vs.
pressure drop relationships.
Figure 5. Flow Curve Showing Cv and K
m
CHOKED FLOW
PLOT OF EQUATION (1)
P1 = CONSTANT
∆P (ALLOWABLE)
C
v
Q
(GPM)
K
m
ΔP
Figure 6. Relationship Between Actual ∆P and ∆P Allowable
∆P (ALLOWABLE)
ACTUAL ∆P
PREDICTED FLOW USING
ACTUAL ∆P
C
v
Q
(GPM)
ACTUAL
FLOW
ΔP
Page 60
Valve Sizing Calculations (Traditional Method)
630
Te c h n i c a l
Use the following equation to determine maximum allowable
pressure drop that is effective in producing ow. Keep in mind,
however, that the limitation on the sizing pressure drop, ∆P
allow
,
does not imply a maximum pressure drop that may be controlled y
the valve.
∆P
allow
= Km (P1 - rc P v) (8)
where:
∆P
allow
= maximum allowable differential pressure for sizing
purposes, psi
Km = valve recovery coefcient from manufacturer’s literature
P1 = body inlet pressure, psia
r
c
= critical pressure ratio determined from Figures 7 and 8
Pv = vapor pressure of the liquid at body inlet temperature,
psia (vapor pressures and critical pressures for
many common liquids are provided in the Physical
Constants of Hydrocarbons and Physical Constants
of Fluids tables; refer to the Table of Contents for the
page number).
After calculating ∆P
allow
, substitute it into the basic liquid sizing
equation
Q = CV ∆P / G
to determine either Q or Cv. If the
actual ∆P is less the ∆P
allow
, then the actual ∆P should be used in
the equation.
The equation used to determine ∆P
allow
should also be used to
calculate the valve body differential pressure at which signicant
cavitation can occur. Minor cavitation will occur at a slightly lower
pressure differential than that predicted by the equation, but should
produce negligible damage in most globe-style control valves.
Consequently, initial cavitation and choked ow occur nearly
simultaneously in globe-style or low-recovery valves.
However, in high-recovery valves such as ball or buttery valves,
signicant cavitation can occur at pressure drops below that which
produces choked ow. So although ∆P
allow
and Km are useful in
predicting choked ow capacity, a separate cavitation index (Kc) is
needed to determine the pressure drop at which cavitation damage
will begin (∆Pc) in high-recovery valves.
The equation can e expressed:
∆PC = KC (P1 - PV) (9)
This equation can be used anytime outlet pressure is greater than
the vapor pressure of the liquid.
Addition of anti-cavitation trim tends to increase the value of Km.
In other words, choked ow and incipient cavitation will occur at
substantially higher pressure drops than was the case without the
anti-cavitation accessory.
Figure 7. Critical Pressure Ratios for Water Figure 8. Critical Pressure Ratios for Liquid Other than Water
USE THIS CURVE FOR WATER. ENTER ON THE ABSCISSA AT THE WATER VAPOR PRESSURE AT THE
VALVE INLET. PROCEED VERTICALLY TO INTERSECT THE
CURVE. MOVE HORIZONTALLY TO THE LEFT TO READ THE CRITICAL
PRESSURE RATIO, RC, ON THE ORDINATE.
USE THIS CURVE FOR LIQUIDS OTHER THAN WATER. DETERMINE THE VAPOR
PRESSURE/CRITICAL PRESSURE RATIO BY DIVIDING THE LIQUID VAPOR PRESSURE
AT THE VALVE INLET BY THE CRITICAL PRESSURE OF THE LIQUID. ENTER ON THE ABSCISSA AT THE
RATIO JUST CALCULATED AND PROCEED VERTICALLY TO
INTERSECT THE CURVE. MOVE HORIZONTALLY TO THE LEFT AND READ THE CRITICAL
PRESSURE RATIO, RC, ON THE ORDINATE.
1.0
0.9
0.8
0.7
0.6
0.5
CRITICAL PRESSURE RATIO—r
c
0 0.20 0.40 0.60 0.80 1.0
VAPOR PRESSURE, PSIA
CRITICAL PRESSURE, PSIA
1.0
0.9
0.8
0.7
0.6
0.5
CRITICAL PRESSURE RATIO—r
c
0 500 1000 1500 2000 2500 3000 3500
VAPOR PRESSURE, PSIA
Page 61
631
Te c h n i c a l
Valve Sizing Calculations (Traditional Method)
Liquid Sizing Summary
The most common use of the basic liquid sizing equation is
to determine the proper valve size for a given set of service
conditions. The rst step is to calculate the required Cv by using
the sizing equation. The ∆P used in the equation must be the actual
valve pressure drop or ∆P
allow
, whichever is smaller. The second
step is to select a valve, from the manufacturer’s literature, with a
Cv equal to or greater than the calculated value.
Accurate valve sizing for liquids requires use of the dual
coefcients of Cv and Km. A single coefcient is not sufcient
to describe both the capacity and the recovery characteristics of
the valve. Also, use of the additional cavitation index factor Kc
is appropriate in sizing high recovery valves, which may develop
damaging cavitation at pressure drops well below the level of the
choked ow.
Liquid Sizing Nomenclature
Cv = valve sizing coefcient for liquid determined
experimentally for each size and style of valve, using
water at standard conditions as the test uid
Cvc = calculated Cv coefcient including correction
for viscosity
Cvr = corrected sizing coefcient required for
viscous applications
∆P = differential pressure, psi
∆P
allow
= maximum allowable differential pressure for sizing
purposes, psi
∆Pc = pressure differential at which cavitation damage
begins, psi
Fv = viscosity correction factor
G = specic gravity of uid (water at 60°F = 1.0000)
Kc = dimensionless cavitation index used in
determining ∆P
c
Km = valve recovery coefcient from
manufacturer’s literature
P1 = body inlet pressure, psia
Pv = vapor pressure of liquid at body inlet
temperature, psia
Q = ow rate capacity, gallons per minute
Q
max
= designation for maximum ow rate, assuming no
viscosity correction required, gallons per minute
Q
pred
= predicted ow rate after incorporating viscosity
correction, gallons per minute
rc = critical pressure ratio
Liquid Sizing Equation Application
EQUATION APPLICATION
1
Basic liquid sizing equation. Use to determine proper valve size for a given set of service conditions.
(Remember that viscosity effects and valve recovery capabilities are not considered in this basic equation.)
2 Use to calculate expected Cv for valve controlling water or other liquids that behave like water.
3
C
vr
= FV C
V
Use to nd actual required Cv for equation (2) after including viscosity correction factor.
4 Use to nd maximum ow rate assuming no viscosity correction is necessary.
5 Use to predict actual ow rate based on equation (4) and viscosity factor correction.
6 Use to calculate corrected sizing coefcient for use in equation (7).
7
ΔP
pred
= G (Q/Cvc)
2
Use to predict pressure drop for viscous liquids.
8
∆P
allow
= Km (P1 - rc P v)
Use to determine maximum allowable pressure drop that is effective in producing ow.
9
∆PC = KC (P1 - PV)
Use to predict pressure drop at which cavitation will begin in a valve with high recovery characteristics.
Q = Cv ΔP / G
CV = Q
G
∆P
Q
max
= Cvr ΔP / G
Q
pred
=
Q
max
F
V
CVC =
C
vr
F
v
Page 62
Valve Sizing Calculations (Traditional Method)
632
Te c h n i c a l
Sizing for Gas or Steam Service
A sizing procedure for gases can be established based on adaptions
of the basic liquid sizing equation. By introducing conversion
factors to change ow units from gallons per minute to cubic
feet per hour and to relate specic gravity in meaningful terms
of pressure, an equation can be derived for the ow of air at
60°F. Because 60°F corresponds to 520° on the Rankine absolute
temperature scale, and because the specic gravity of air at 60°F
is 1.0, an additional factor can be included to compare air at 60°F
with specic gravity (G) and absolute temperature (T) of any other
gas. The resulting equation an be written:
(A)
The equation shown above, while valid at very low pressure
drop ratios, has been found to be very misleading when the ratio
of pressure drop (∆P) to inlet pressure (P1) exceeds 0.02. The
deviation of actual ow capacity from the calculated ow capacity
is indicated in Figure 8 and results from compressibility effects and
critical ow limitations at increased pressure drops.
Critical ow limitation is the more signicant of the two problems
mentioned. Critical ow is a choked ow condition caused by
increased gas velocity at the vena contracta. When velocity at the
vena contracta reaches sonic velocity, additional increases in ∆P
by reducing downstream pressure produce no increase in ow.
So, after critical ow condition is reached (whether at a pressure
drop/inlet pressure ratio of about 0.5 for glove valves or at much
lower ratios for high recovery valves) the equation above becomes
completely useless. If applied, the Cv equation gives a much higher
indicated capacity than actually will exist. And in the case of a
high recovery valve which reaches critical ow at a low pressure
drop ratio (as indicated in Figure 8), the critical ow capacity of
the valve may be over-estimated by as much as 300 percent.
The problems in predicting critical ow with a Cv-based equation
led to a separate gas sizing coefcient based on air ow tests.
The coefcient (Cg) was developed experimentally for each
type and size of valve to relate critical ow to absolute inlet
pressure. By including the correction factor used in the previous
equation to compare air at 60°F with other gases at other absolute
temperatures, the critical ow equation an be written:
Q
critical
= CgP1 520 / GT (B)
Figure 9. Critical Flow for High and Low Recovery
Valves with Equal C
v
Universal Gas Sizing Equation
To account for differences in ow geometry among valves,
equations (A) and (B) were consolidated by the introduction of
an additional factor (C1). C1 is dened as the ratio of the gas
sizing coefcient and the liquid sizing coefcient and provides a
numerical indicator of the valve’s recovery capabilities. In general,
C1 values can range from about 16 to 37, based on the individual
valve’s recovery characteristics. As shown in the example, two
valves with identical ow areas and identical critical ow (Cg)
capacities can have widely differing C1 values dependent on the
effect internal ow geometry has on liquid ow capacity through
each valve. Example:
High Recovery Valve
Cg = 4680
Cv = 254
C1 = Cg/Cv
= 4680/254
= 18.4
Low Recovery Valve
Cg = 4680
Cv = 135
C1 = Cg/C
v
= 4680/135
= 34.7
Q
SCFH
= 59.64 CVP
1
∆P
P
1
520
GT
HIGH RECOVERY
LOW RECOVERY
C
v
Q
∆P
P
1
= 0.5
∆P
P
1
= 0.15
∆P / P
1
Page 63
633
Te c h n i c a l
Valve Sizing Calculations (Traditional Method)
So we see that two sizing coefcients are needed to accurately
size valves for gas ow —Cg to predict ow based on physical size
or ow area, and C1 to account for differences in valve recovery
characteristics. A blending equation, called the Universal Gas
Sizing Equation, combines equations (A) and (B) by means of a
sinusoidal function, and is based on the “perfect gas” laws. It can
be expressed in either of the following manners:
(C)
OR
(D)
In either form, the equation indicates critical ow when the sine
function of the angle designated within the brackets equals unity.
The pressure drop ratio at which critical ow occurs is known
as the critical pressure drop ratio. It occurs when the sine angle
reaches π/2 radians in equation (C) or 90 degrees in equation
(D). As pressure drop across the valve increases, the sine angle
increases from zero up to π/2 radians (90°). If the angle were
allowed to increase further, the equations would predict a decrease
in ow. Because this is not a realistic situation, the angle must be
limited to 90 degrees maximum.
Although “perfect gases,” as such, do not exist in nature, there are a
great many applications where the Universal Gas Sizing Equation,
(C) or (D), provides a very useful and usable approximation.
General Adaptation for Steam and Vapors
The density form of the Universal Gas Sizing Equation is the most
general form and can be used for both perfect and non-perfect gas
applications. Applying the equation requires knowledge of one
additional condition not included in previous equations, that being
the inlet gas, steam, or vapor density (d1) in pounds per cubic foot.
(Steam density can be determined from tables.)
Then the following adaptation of the Universal Gas Sizing
Equation can be applied:
(E)
Special Equation Form for Steam Below 1000 psig
If steam applications do not exceed 1000 psig, density changes can
be compensated for by using a special adaptation of the Universal
Gas Sizing Equation. It incorporates a factor for amount of
superheat in degrees Fahrenheit (Tsh) and also a sizing coefcient
(Cs) for steam. Equation (F) eliminates the need for nding the
density of superheated steam, which was required in Equation
(E). At pressures below 1000 psig, a constant relationship exists
between the gas sizing coefcient (Cg) and the steam coefcient
(Cs). This relationship can be expressed: Cs = Cg/20. For higher
steam pressure application, use Equation (E).
(F)
Gas and Steam Sizing Summary
The Universal Gas Sizing Equation can be used to determine
the ow of gas through any style of valve. Absolute units of
temperature and pressure must be used in the equation. When the
critical pressure drop ratio causes the sine angle to be 90 degrees,
the equation will predict the value of the critical ow. For service
conditions that would result in an angle of greater than 90 degrees,
the equation must be limited to 90 degrees in order to accurately
determine the critical ow.
Most commonly, the Universal Gas Sizing Equation is used to
determine proper valve size for a given set of service conditions.
The rst step is to calculate the required Cg by using the Universal
Gas Sizing Equation. The second step is to select a valve from
the manufacturer’s literature. The valve selected should have a C
g
which equals or exceeds the calculated value. Be certain that the
assumed C1 value for the valve is selected from the literature.
It is apparent that accurate valve sizing for gases that requires use
of the dual coefcient is not sufcient to describe both the capacity
and the recovery characteristics of the valve.
Proper selection of a control valve for gas service is a highly
technical problem with many factors to be considered. Leading
valve manufacturers provide technical information, test data, sizing
catalogs, nomographs, sizing slide rules, and computer or calculator
programs that make valve sizing a simple and accurate procedure.
Q
lb/hr
=
CS P
1
1 + 0.00065T
sh
SIN
3417
C
1
∆P
P
1
Deg
Q
SCFH
=
520
GT
Cg P1 SIN
59.64
C
1
∆P
P
1
rad
Q
SCFH
=
520
GT
Cg P1 SIN
3417
C
1
∆P
P
1
Deg
Q
lb/hr
= 1.06 d1 P1 Cg SIN Deg
3417
C
1
∆P
P
1
Page 64
Valve Sizing Calculations (Traditional Method)
634
Te c h n i c a l
Gas and Steam Sizing Equation Application
EQUATION APPLICATION
A Use only at very low pressure drop (DP/P1) ratios of 0.02 or less.
B Use only to determine critical ow capacity at a given inlet pressure.
C
D
OR
Universal Gas Sizing Equation.
Use to predict ow for either high or low recovery valves, for any gas adhering to the
perfect gas laws, and under any service conditions.
E
Use to predict ow for perfect or non-perfect gas sizing applications, for any vapor
including steam, at any service condition when uid density is known.
F Use only to determine steam ow when inlet pressure is 1000 psig or less.
C
1
= Cg/C
v
C
g
= gas sizing coefcient
C
s
= steam sizing coefcient, Cg/20
C
v
= liquid sizing coefcient
d
1
= density of steam or vapor at inlet, pounds/cu. foot
G = gas specic gravity (air = 1.0)
P
1
= valve inlet pressure, psia
Gas and Steam Sizing Nomenclature
∆P = pressure drop across valve, psi
Q
critical
= critical ow rate, SCFH
Q
SCFH
= gas ow rate, SCFH
Q
lb/hr
= steam or vapor ow rate, pounds per hour
T = absolute temperature of gas at inlet, degrees Rankine
T
sh
= degrees of superheat, °F
Q
SCFH
= 59.64 CVP
1
∆P
P
1
520
GT
Q
critical
= CgP1 520 / GT
Q
lb/hr
=
CS P
1
1 + 0.00065T
sh
SIN
3417
C
1
∆P
P
1
Deg
Q
SCFH
=
520
GT
Cg P1 SIN
59.64
C
1
∆P
P
1
rad
Q
SCFH
=
520
GT
Cg P1 SIN
3417
C
1
∆P
P
1
Deg
Q
lb/hr
= 1.06 d1 P1 Cg SIN Deg
3417
C
1
∆P
P
1
Page 65
Valve Sizing (Standardized Method)
635
Te c h n i c a l
Introduction
Fisher® regulators and valves have traditionally been sized using
equations derived by the company. There are now standardized
calculations that are becoming accepted world wide. Some product
literature continues to demonstrate the traditional method, but the
trend is to adopt the standardized method. Therefore, both methods
are covered in this application guide.
Liquid Valve Sizing
Standardization activities for control valve sizing can be traced
back to the early 1960s when a trade association, the Fluids Control
Institute, published sizing equations for use with both compressible
and incompressible uids. The range of service conditions that
could be accommodated accurately by these equations was quite
narrow, and the standard did not achieve a high degree of acceptance.
In 1967, the ISA established a committee to develop and publish
standard equations. The efforts of this committee culminated in
a valve sizing procedure that has achieved the status of American
National Standard. Later, a committee of the International
Electrotechnical Commission (IEC) used the ISA works as a basis to
formulate international standards for sizing control valves. (Some
information in this introductory material has been extracted from
ANSI/ISA S75.01 standard with the permission of the publisher,
the ISA.) Except for some slight differences in nomenclature and
procedures, the ISA and IEC standards have been harmonized.
ANSI/ISA Standard S75.01 is harmonized with IEC Standards 5342-1 and 534-2-2. (IEC Publications 534-2, Sections One and Two for
incompressible and compressible uids, respectively.)
In the following sections, the nomenclature and procedures are explained,
and sample problems are solved to illustrate their use.
Sizing Valves for Liquids
Following is a step-by-step procedure for the sizing of control valves for
liquid ow using the IEC procedure. Each of these steps is important
and must be considered during any valve sizing procedure. Steps 3 and
4 concern the determination of certain sizing factors that may or may not
be required in the sizing equation depending on the service conditions of
the sizing problem. If one, two, or all three of these sizing factors are to
be included in the equation for a particular sizing problem, refer to the
appropriate factor determination section(s) located in the text after the
sixth step.
1. Specify the variables required to size the valve as follows:
• Desired design
• Process uid (water, oil, etc.), and
• Appropriate service conditions q or w, P1, P2, or ∆P, T1, Gf, Pv,
Pc, and υ.
The ability to recognize which terms are appropriate for
a specic sizing procedure can only be acquired through
experience with different valve sizing problems. If any
of the above terms appears to be new or unfamiliar, refer
to the Abbreviations and Terminology Table 3-1 for a
complete denition.
2. Determine the equation constant, N.
N is a numerical constant contained in each of the ow equations
to provide a means for using different systems of units. Values
for these various constants and their applicable units are given in
the Equation Constants Table 3-2.
Use N1, if sizing the valve for a ow rate in volumetric units
(GPM or Nm3/h).
Use N6, if sizing the valve for a ow rate in mass units
(pound/hr or kg/hr).
3. Determine Fp, the piping geometry factor.
Fp is a correction factor that accounts for pressure losses due
to piping ttings such as reducers, elbows, or tees that might
be attached directly to the inlet and outlet connections of the
control valve to be sized. If such ttings are attached to the
valve, the Fp factor must be considered in the sizing procedure.
If, however, no ttings are attached to the valve, Fp has a value of
1.0 and simply drops out of the sizing equation.
For rotary valves with reducers (swaged installations),
and other valve designs and tting styles, determine the Fp
factors by using the procedure for determining Fp, the Piping
Geometry Factor, page 637.
4. Determine q
max
(the maximum ow rate at given upstream
conditions) or ∆P
max
(the allowable sizing pressure drop).
The maximum or limiting ow rate (q
max
), commonly called
choked ow, is manifested by no additional increase in ow
rate with increasing pressure differential with xed upstream
conditions. In liquids, choking occurs as a result of
vaporization of the liquid when the static pressure within
the valve drops below the vapor pressure of the liquid.
The IEC standard requires the calculation of an allowable
sizing pressure drop (∆P
max
), to account for the possibility
of choked ow conditions within the valve. The calculated
∆P
max
value is compared with the actual pressure drop specied
in the service conditions, and the lesser of these two values
is used in the sizing equation. If it is desired to use ∆P
max
to
account for the possibility of choked ow conditions, it can
be calculated using the procedure for determining q
max
, the
Maximum Flow Rate, or ∆P
max
, the Allowable Sizing Pressure
Drop. If it can be recognized that choked ow conditions will
not develop within the valve, ∆P
max
need not be calculated.
5. Solve for required C
v
, using the appropriate equation:
• For volumetric ow rate units:
Cv =
q
N
1Fp
P1 - P
2
G
f
• For mass ow rate units:
Cv=
w
N6F
p
(P1-P2) γ
In addition to Cv, two other ow coefcients, Kv and Av, are
used, particularly outside of North America. The following
relationships exist:
Kv= (0.865) (Cv)
Av= (2.40 x 10-5) (Cv)
6. Select the valve size using the appropriate ow coefcient
table and the calculated Cv value.
Page 66
Valve Sizing (Standardized Method)
636
Te c h n i c a l
Table 3-1. Abbreviations and Terminology
SYMBOL SYMBOL
C
v
Valve sizing coefcient P
1
Upstream absolute static pressure
d Nominal valve size P
2
Downstream absolute static pressure
D Internal diameter of the piping P
c
Absolute thermodynamic critical pressure
F
d
Valve style modier, dimensionless P
v
Vapor pressure absolute of liquid
at inlet temperature
F
F
Liquid critical pressure ratio factor,
dimensionless
∆P Pressure drop (P1-P2) across the valve
F
k
Ratio of specic heats factor, dimensionless ∆P
max(L)
Maximum allowable liquid sizing
pressure drop
F
L
Rated liquid pressure recovery factor,
dimensionless
∆P
max(LP)
Maximum allowable sizing pressure
drop with attached ttings
F
LP
Combined liquid pressure recovery factor and
piping geometry factor of valve with attached
ttings (when there are no attached ttings, FLP
equals FL), dimensionless
q Volume rate of ow
F
P
Piping geometry factor, dimensionless q
max
Maximum ow rate (choked ow conditions)
at given upstream conditions
G
f
Liquid specic gravity (ratio of density of liquid at
owing temperature to density of water at 60°F),
dimensionless
T
1
Absolute upstream temperature
(deg Kelvin or deg Rankine)
G
g
Gas specic gravity (ratio of density of owing
gas to density of air with both at standard
conditions
(1)
, i.e., ratio of molecular weight of gas
to molecular weight of air), dimensionless
w Mass rate of ow
k Ratio of specic heats, dimensionless x
Ratio of pressure drop to upstream absolute
static pressure (∆P/P1), dimensionless
K Head loss coefcient of a device, dimensionless x
T
Rated pressure drop ratio factor, dimensionless
M Molecular weight, dimensionless Y
Expansion factor (ratio of ow coefcient for a
gas to that for a liquid at the same Reynolds
number), dimensionless
N Numerical constant
Z Compressibility factor, dimensionless
γ
1
Specic weight at inlet conditions
υ
Kinematic viscosity, centistokes
1. Standard conditions are dened as 60°F and 14.7 psia.
Table 3-2. Equation Constants
(1)
N w q p
(2)
γ
T d, D
N
1
0.0865
0.865
1.00
- - - -
- - - -
- - - -
Nm3/h
Nm3/h
GPM
kPa
bar
psia
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
N
2
0.00214
890
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
mm
inch
N
5
0.00241
1000
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
mm
inch
N
6
2.73
27.3
63.3
kg/hr
kg/hr
pound/hr
- - - -
- - - -
- - - -
kPa
bar
psia
kg/m
3
kg/m
3
pound/ft
3
- - - -
- - - -
- - - -
- - - -
- - - -
- - - -
N
7
(3)
Normal Conditions
TN = 0°C
3.94
394
- - - -
- - - -
Nm3/h
Nm3/h
kPa
bar
- - - -
- - - -
deg Kelvin
deg Kelvin
- - - -
- - - -
Standard Conditions
Ts = 16°C
4.17
417
- - - -
- - - -
Nm3/h
Nm3/h
kPa
bar
- - - -
- - - -
deg Kelvin
deg Kelvin
- - - -
- - - -
Standard Conditions
Ts = 60°F
1360 - - - - SCFH psia - - - - deg Rankine - - - -
N
8
0.948
94.8
19.3
kg/hr
kg/hr
pound/hr
- - - -
- - - -
- - - -
kPa
bar
psia
- - - -
- - - -
- - - -
deg Kelvin
deg Kelvin
deg Rankine
- - - -
- - - -
- - - -
N
9
(3)
Normal Conditions
TN = 0°C
21.2
2120
- - - -
- - - -
Nm3/h
Nm3/h
kPa
bar
- - - -
- - - -
deg Kelvin
deg Kelvin
- - - -
- - - -
Standard Conditions
TS = 16°C
22.4
2240
- - - -
- - - -
Nm3/h
Nm3/h
kPa
bar
- - - -
- - - -
deg Kelvin
deg Kelvin
- - - -
- - - -
Standard Conditions
TS = 60°F
7320 - - - - SCFH psia - - - - deg Rankine - - - -
1. Many of the equations used in these sizing procedures contain a numerical constant, N, along with a numerical subscript. These numerical constants provide a means for using
different units in the equations. Values for the various constants and the applicable units are given in the above table. For example, if the ow rate is given in U.S. GPM and
the pressures are psia, N1 has a value of 1.00. If the ow rate is Nm3/h and the pressures are kPa, the N1 constant becomes 0.0865.
2. All pressures are absolute.
3. Pressure base is 101.3 kPa (1,01 bar) (14.7 psia).
Page 67
Valve Sizing (Standardized Method)
637
Te c h n i c a l
Determining Piping Geometry Factor (Fp)
Determine an Fp factor if any ttings such as reducers, elbows, or
tees will be directly attached to the inlet and outlet connections
of the control valve that is to be sized. When possible, it is
recommended that Fp factors be determined experimentally by
using the specied valve in actual tests.
Calculate the Fp factor using the following equation:
Fp=
1 +
∑K
N
2
C
v
d
2
2
-1/2
where,
N2 = Numerical constant found in the Equation Constants table
d = Assumed nominal valve size
Cv = Valve sizing coefcient at 100% travel for the assumed
valve size
In the above equation, the ∑K term is the algebraic sum of the
velocity head loss coefcients of all of the ttings that are attached to
the control valve.
∑K = K1 + K2 + KB1 - K
B2
where,
K1 = Resistance coefcient of upstream ttings
K2 = Resistance coefcient of downstream ttings
KB1 = Inlet Bernoulli coefcient
KB2 = Outlet Bernoulli coefcient
The Bernoulli coefcients, KB1 and KB2, are used only when the
diameter of the piping approaching the valve is different from the
diameter of the piping leaving the valve, whereby:
KB1 or K
B2
= 1-
d
D
where,
d = Nominal valve size
D = Internal diameter of piping
If the inlet and outlet piping are of equal size, then the Bernoulli
coefcients are also equal, KB1 = KB2, and therefore they are dropped
from the equation.
The most commonly used tting in control valve installations is
the short-length concentric reducer. The equations for this tting
are as follows:
• For an inlet reducer:
K1= 0.5
1-
d
2
D
2
2
4
• For an outlet reducer:
K2= 1.0
1-
d
2
D
2
2
• For a valve installed between identical reducers:
K1 + K2= 1.5
1-
d
2
D
2
2
Determining Maximum Flow Rate (q
max
)
Determine either q
max
or ∆P
max
if it is possible for choked ow to
develop within the control valve that is to be sized. The values can
be determined by using the following procedures.
q
max
=
N
1FLCV
P1 - FF P
V
G
f
Values for FF, the liquid critical pressure ratio factor, can be
obtained from Figure 3-1, or from the following equation:
FF = 0.96 - 0.28
P
V
P
C
Values of FL, the recovery factor for rotary valves installed without
ttings attached, can be found in published coefcient tables. If the
given valve is to be installed with ttings such as reducer attached to
it, FL in the equation must be replaced by the quotient FLP/FP, where:
FLP =
K
1
N
2
C
V
d
2
2
+
1
F
L
2
-1/2
and
K1 = K1 + K
B1
where,
K1 = Resistance coefcient of upstream ttings
KB1 = Inlet Bernoulli coefcient
(See the procedure for Determining Fp, the Piping Geometry Factor,
for denitions of the other constants and coefcients used in the
above equations.)
Page 68
Valve Sizing (Standardized Method)
638
Te c h n i c a l
LIQUID CRITICAL PRESSURE RATIO
FACTOR—F
F
ABSOLUTE VAPOR PRESSURE-bar
ABSOLUTE VAPOR PRESSURE-PSIA
USE THIS CURVE FOR WATER, ENTER ON THE ABSCISSA AT THE WATER VAPOR
PRESSURE AT THE VALVE INLET, PROCEED VERTICALLY TO INTERSECT THE CURVE,
MOVE HORIZONTALLY TO THE LEFT TO READ THE CRITICAL PRESSURE RATIO, FF, ON
THE ORDINATE.
0
500
1000 1500
2000
2500 3000
3500
0.5
0.6
0.7
0.8
0.9
1.0
34
69
103
138
172 207
241
A2737-1
Figure 3-1. Liquid Critical Pressure Ratio Factor for Water
Determining Allowable Sizing Pressure Drop (∆P
max
)
∆P
max
(the allowable sizing pressure drop) can be determined from
the following relationships:
For valves installed without ttings:
∆P
max(L)
= F
L
2
(P1 - FF PV)
For valves installed with ttings attached:
∆P
max(LP)
=
F
LP
F
P
2
(P1 - FF PV)
where,
P1 = Upstream absolute static pressure
P2 = Downstream absolute static pressure
Pv = Absolute vapor pressure at inlet temperature
Values of FF, the liquid critical pressure ratio factor, can be obtained
from Figure 3-1 or from the following equation:
FF = 0.96 - 0.28
P
V
P
c
An explanation of how to calculate values of FLP, the recovery
factor for valves installed with ttings attached, is presented in the
preceding procedure Determining q
max
(the Maximum Flow Rate).
Once the ∆P
max
value has been obtained from the appropriate
equation, it should be compared with the actual service pressure
differential (∆P = P1 - P2). If ∆P
max
is less than ∆P, this is an
indication that choked ow conditions will exist under the service
conditions specied. If choked ow conditions do exist (∆P
max
< P1 - P2), then step 5 of the procedure for Sizing Valves for
Liquids must be modied by replacing the actual service pressure
differential (P1 - P2) in the appropriate valve sizing equation with
the calculated ∆P
max
value.
Note
Once it is known that choked ow conditions will
develop within the specied valve design (∆P
max
is
calculated to be less than ∆P), a further distinction
can be made to determine whether the choked
ow is caused by cavitation or ashing. The
choked ow conditions are caused by ashing if
the outlet pressure of the given valve is less than
the vapor pressure of the owing liquid. The
choked ow conditions are caused by cavitation if
the outlet pressure of the valve is greater than the
vapor pressure of the owing liquid.
Liquid Sizing Sample Problem
Assume an installation that, at initial plant startup, will not be
operating at maximum design capability. The lines are sized
for the ultimate system capacity, but there is a desire to install a
control valve now which is sized only for currently anticipated
requirements. The line size is 8-inch (DN 200) and an ASME
CL300 globe valve with an equal percentage cage has been
specied. Standard concentric reducers will be used to install the
valve into the line. Determine the appropriate valve size.
Page 69
Valve Sizing (Standardized Method)
639
Te c h n i c a l
1. Specify the necessary variables required to size the valve:
• Desired Valve Design—ASME CL300 globe valve with
equal percentage cage and an assumed valve size of 3-inches.
• Process Fluid—liquid propane
• Service Conditions—q = 800 GPM (3028 l/min)
P1 = 300 psig (20,7 bar) = 314.7 psia (21,7 bar a)
P2 = 275 psig (19,0 bar) = 289.7 psia (20,0 bar a)
∆P = 25 psi (1,7 bar)
T1 = 70°F (21°C)
Gf = 0.50
Pv = 124.3 psia (8,6 bar a)
Pc = 616.3 psia (42,5 bar a)
2. Use an N1 value of 1.0 from the Equation Constants table.
3. Determine Fp, the piping geometry factor.
Because it is proposed to install a 3-inch valve in
an 8-inch (DN 200) line, it will be necessary to determine
the piping geometry factor, Fp, which corrects for losses
caused by ttings attached to the valve.
Fp =
1 +
∑K
N
2
C
v
d
2
2
-1/2
where,
N2 = 890, from the Equation Constants table
d = 3-inch (76 mm), from step 1
Cv = 121, from the ow coefcient table for an ASME CL300,
3-inch globe valve with equal percentage cage
To compute ∑K for a valve installed between identical
concentric reducers:
∑K = K1 + K
2
= 1.5
1 -
d
2
D
2
2
= 1.5
1 -
(3)
2
(8)
2
2
= 1.11
USE THIS CURVE FOR LIQUIDS OTHER THAN WATER. DETERMINE THE VAPOR PRESSURE/
CRITICAL PRESSURE RATIO BY DIVIDING THE LIQUID VAPOR PRESSURE AT THE VALVE INLET
BY THE CRITICAL PRESSURE OF THE LIQUID. ENTER ON THE ABSCISSA AT THE RATIO JUST
CALCULATED AND PROCEED VERTICALLY TO INTERSECT THE CURVE. MOVE HORIZONTALLY
TO THE LEFT AND READ THE CRITICAL PRESSURE RATIO, FF, ON THE ORDINATE.
Figure 3-2. Liquid Critical Pressure Ratio Factor for Liquids Other Than Water
ABSOLUTE VAPOR PRESSURE
ABSOLUTE THERMODYNAMIC CRITICAL PRESSURE
P
v
P
c
0
0.10
0.20 0.30
0.40
0.50 0.60
0.70
0.5
0.6
0.7
0.8
0.9
1.0
0.80
0.90
1.00
LIQUID CRITICAL PRESSURE RATIO
FACTOR—F
F
Page 70
Valve Sizing (Standardized Method)
640
Te c h n i c a l
where,
D = 8-inch (203 mm), the internal diameter of the piping so,
Fp =
1 +
1.11
890
121
3
2
2
-1/2
= 0.90
4. Determine ∆P
max
(the Allowable Sizing Pressure Drop.)
Based on the small required pressure drop, the ow will not
be choked (∆P
max
> ∆P).
5. Solve for Cv, using the appropriate equation.
Cv =
q
N1F
p
P1 - P
2
G
f
=
800
(1.0) (0.90)
25
0.5
= 125.7
6. Select the valve size using the ow coefcient table and the
calculated Cv value.
The required Cv of 125.7 exceeds the capacity of the assumed
valve, which has a Cv of 121. Although for this example it
may be obvious that the next larger size (4-inch) would
be the correct valve size, this may not always be true, and a
repeat of the above procedure should be carried out.
Assuming a 4-inches valve, Cv = 203. This value was
determined from the ow coefcient table for an ASME
CL300, 4-inch globe valve with an equal percentage cage.
Recalculate the required Cv using an assumed Cv value of
203 in the Fp calculation.
where,
∑K = K1 + K
2
= 1.5
1 -
2
d
2
D
2
= 1.5
1 -
2
16
64
= 0.84
and
Fp =
1.0 +
∑K
N
2
C
v
d
2
2
-1/2
=
1.0 +
0.84
890
203
4
2
2
-1/2
= 0.93
and
Cv =
N1F
p
P1 - P
2
G
f
q
=
(1.0) (0.93)
25
0.5
800
= 121.7
This solution indicates only that the 4-inch valve is large enough to
satisfy the service conditions given. There may be cases, however,
where a more accurate prediction of the Cv is required. In such
cases, the required Cv should be redetermined using a new Fp value
based on the Cv value obtained above. In this example, Cv is 121.7,
which leads to the following result:
Fp =
1.0 +
∑K
N
2
C
v
d
2
2
-1/2
=
1.0 +
0.84
890
121.7
4
2
2
-1/2
= 0.97
The required Cv then becomes:
Cv =
N1F
p
P1 - P
2
G
f
q
=
(1.0) (0.97)
25
0.5
800
= 116.2
Because this newly determined Cv is very close to the Cv used
initially for this recalculation (116.2 versus 121.7), the valve sizing
procedure is complete, and the conclusion is that a 4-inch valve
opened to about 75% of total travel should be adequate for the
required specications.
Page 71
Valve Sizing (Standardized Method)
641
Te c h n i c a l
Gas and Steam Valve Sizing
Sizing Valves for Compressible Fluids
Following is a six-step procedure for the sizing of control valves
for compressible ow using the ISA standardized procedure.
Each of these steps is important and must be considered
during any valve sizing procedure. Steps 3 and 4 concern the
determination of certain sizing factors that may or may not
be required in the sizing equation depending on the service
conditions of the sizing problem. If it is necessary for one or
both of these sizing factors to be included in the sizing equation
for a particular sizing problem, refer to the appropriate factor
determination section(s), which is referenced and located in the
following text.
1. Specify the necessary variables required to size the valve
as follows:
• Desired valve design (e.g. balanced globe with linear cage)
• Process uid (air, natural gas, steam, etc.) and
• Appropriate service conditions—
q, or w, P1, P2 or P, T1, Gg, M, k, Z, and γ
1
The ability to recognize which terms are appropriate
for a specic sizing procedure can only be acquired through
experience with different valve sizing problems. If any
of the above terms appear to be new or unfamiliar, refer to
the Abbreviations and Terminology Table 3-1 in Liquid
Valve Sizing Section for a complete denition.
2. Determine the equation constant, N.
N is a numerical constant contained in each of the ow
equations to provide a means for using different systems of
units. Values for these various constants and their applicable
units are given in the Equation Constants Table 3-2 in
Liquid Valve Sizing Section.
Use either N7 or N9 if sizing the valve for a ow rate in
volumetric units (SCFH or Nm3/h). Which of the two
constants to use depends upon the specied service
conditions. N7 can be used only if the specic gravity, Gg,
of the following gas has been specied along with the other
required service conditions. N9 can be used only if the
molecular weight, M, of the gas has been specied.
Use either N6 or N8 if sizing the valve for a ow rate in mass
units (pound/hr or kg/hr). Which of the two constants to use
depends upon the specied service conditions. N6 can
be used only if the specic weight, γ1, of the owing gas has
been specied along with the other required service
conditions. N8 can be used only if the molecular weight, M,
of the gas has been specied.
3. Determine Fp, the piping geometry factor.
Fp is a correction factor that accounts for any pressure losses
due to piping ttings such as reducers, elbows, or tees that
might be attached directly to the inlet and outlet connections
of the control valves to be sized. If such ttings are attached
to the valve, the Fp factor must be considered in the sizing
procedure. If, however, no ttings are attached to the valve, Fp
has a value of 1.0 and simply drops out of the sizing equation.
Also, for rotary valves with reducers and other valve designs
and tting styles, determine the Fp factors by using the procedure
for Determining Fp, the Piping Geometry Factor, which is located
in Liquid Valve Sizing Section.
4. Determine Y, the expansion factor, as follows:
where,
Fk = k/1.4, the ratio of specic heats factor
k = Ratio of specic heats
x = P/P1, the pressure drop ratio
xT = The pressure drop ratio factor for valves installed without
attached ttings. More denitively, xT is the pressure
drop ratio required to produce critical, or maximum, ow
through the valve when Fk = 1.0
If the control valve to be installed has ttings such as reducers
or elbows attached to it, then their effect is accounted for in the
expansion factor equation by replacing the xT term with a new
factor xTP. A procedure for determining the xTP factor is
described in the following section for Determining xTP, the
Pressure Drop Ratio Factor.
Note
Conditions of critical pressure drop are realized
when the value of x becomes equal to or exceeds the
appropriate value of the product of either Fk xT or
Fk xTP at which point:
Although in actual service, pressure drop ratios can, and often
will, exceed the indicated critical values, this is the point where
critical ow conditions develop. Thus, for a constant P1,
decreasing P2 (i.e., increasing P) will not result in an increase in
the ow rate through the valve. Values of x, therefore, greater
than the product of either FkxT or FkxTP must never be substituted
in the expression for Y. This means that Y can never be less
than 0.667. This same limit on values of x also applies to the ow
equations that are introduced in the next section.
5. Solve for the required Cv using the appropriate equation:
For volumetric ow rate units—
• If the specic gravity, Gg, of the gas has been specied:
y = 1 - = 1 - 1/3 = 0.667
x
3Fk x
T
Y = 1 -
x
3Fk x
T
Cv =
q
N7 FP P1 Y
Gg T1 Z
x
Page 72
Valve Sizing (Standardized Method)
642
Te c h n i c a l
• If the molecular weight, M, of the gas has been specied:
For mass ow rate units—
• If the specic weight, γ1, of the gas has been specied:
• If the molecular weight, M, of the gas has been specied:
In addition to Cv, two other ow coefcients, Kv and Av, are
used, particularly outside of North America. The following
relationships exist:
Kv = (0.865)(Cv)
Av = (2.40 x 10-5)(Cv)
6. Select the valve size using the appropriate ow coefcient table
and the calculated Cv value.
Determining xTP, the Pressure Drop Ratio Factor
If the control valve is to be installed with attached ttings such as
reducers or elbows, then their effect is accounted for in the expansion
factor equation by replacing the xT term with a new factor, xTP.
where,
N5 = Numerical constant found in the Equation Constants table
d = Assumed nominal valve size
Cv = Valve sizing coefcient from ow coefcient table at
100% travel for the assumed valve size
Fp = Piping geometry factor
xT = Pressure drop ratio for valves installed without ttings attached.
xT values are included in the ow coefcient tables
In the above equation, Ki, is the inlet head loss coefcient, which is
dened as:
Ki = K1 + K
B1
where,
K1 = Resistance coefcient of upstream ttings (see the procedure
for Determining Fp, the Piping Geometry Factor, which is
contained in the section for Sizing Valves for Liquids).
KB1 = Inlet Bernoulli coefcient (see the procedure for Determining
Fp, the Piping Geometry Factor, which is contained in the
section for Sizing Valves for Liquids).
Compressible Fluid Sizing Sample Problem No. 1
Determine the size and percent opening for a Fisher® Design V250
ball valve operating with the following service conditions. Assume
that the valve and line size are equal.
1. Specify the necessary variables required to size the valve:
• Desired valve design—Design V250 valve
• Process uid—Natural gas
• Service conditions—
P1 = 200 psig (13,8 bar) = 214.7 psia (14,8 bar)
P2 = 50 psig (3,4 bar) = 64.7 psia (4,5 bar)
P = 150 psi (10,3 bar)
x = P/P1 = 150/214.7 = 0.70
T1 = 60°F (16°C) = 520°R
M = 17.38
Gg = 0.60
k = 1.31
q = 6.0 x 106 SCFH
2. Determine the appropriate equation constant, N, from the
Equation Constants Table 3-2 in Liquid Valve Sizing Section.
Because both Gg and M have been given in the service
conditions, it is possible to use an equation containing either N
7
or N9. In either case, the end result will be the same. Assume
that the equation containing Gg has been arbitrarily selected for
this problem. Therefore, N7 = 1360.
3. Determine Fp, the piping geometry factor.
Since valve and line size are assumed equal, Fp = 1.0.
4. Determine Y, the expansion factor.
Fk =
=
= 0.94
It is assumed that an 8-inch Design V250 valve will be adequate
for the specied service conditions. From the ow coefcient
Table 4-2, xT for an 8-inch Design V250 valve at 100% travel
is 0.137.
x = 0.70 (This was calculated in step 1.)
Cv =
q
N7 FP P1 Y
M T1 Z
x
Cv =
w
N6 FP Y
x P1 γ
1
Cv =
w
N8 FP P1 Y
T1 Z
x M
k
1.40
1.40
1.31
xTP = 1 +
xT
F
p
2
xT K
i
N
5
Cv
d2
2
-1
Page 73
Valve Sizing (Standardized Method)
643
Te c h n i c a l
Since conditions of critical pressure drop are realized
when the calculated value of x becomes equal to or
exceeds the appropriate value of FkxT, these values
should be compared.
FkxT = (0.94) (0.137)
= 0.129
Because the pressure drop ratio, x = 0.70 exceeds the
calculated critical value, FkxT = 0.129, choked ow
conditions are indicated. Therefore, Y = 0.667, and
x = FkxT = 0.129.
5. Solve for required Cv using the appropriate equation.
The compressibility factor, Z, can be assumed to be 1.0 for
the gas pressure and temperature given and Fp = 1 because
valve size and line size are equal.
So,
6. Select the valve size using the ow coefcient table and
the calculated Cv value.
The above result indicates that the valve is adequately sized
(rated Cv = 2190). To determine the percent valve opening,
note that the required Cv occurs at approximately 83 degrees
for the 8-inch Design V250 valve. Note also that, at
83 degrees opening, the xT value is 0.252, which is
substantially different from the rated value of 0.137 used
initially in the problem. The next step is to rework the
problem using the xT value for 83 degrees travel.
The FkxT product must now be recalculated.
x = Fk x
T
= (0.94) (0.252)
= 0.237
The required Cv now becomes:
The reason that the required Cv has dropped so dramatically
is attributable solely to the difference in the xT values at rated
and 83 degrees travel. A Cv of 1118 occurs between 75 and
80 degrees travel.
The appropriate ow coefcient table indicates that xT is higher at
75 degrees travel than at 80 degrees travel. Therefore, if the
problem were to be reworked using a higher xT value, this should
result in a further decline in the calculated required Cv.
Reworking the problem using the xT value corresponding to
78 degrees travel (i.e., xT = 0.328) leaves:
x = Fk x
T
= (0.94) (0.328)
= 0.308
and,
The above Cv of 980 is quite close to the 75 degree travel Cv. The
problem could be reworked further to obtain a more precise predicted
opening; however, for the service conditions given, an 8-inch
Design V250 valve installed in an 8-inch (203 mm) line
will be approximately 75 degrees open.
Compressible Fluid Sizing Sample Problem No. 2
Assume steam is to be supplied to a process designed to operate at
250 psig (17 bar). The supply source is a header maintained at
500 psig (34,5 bar) and 500°F (260°C). A 6-inch (DN 150) line
from the steam main to the process is being planned. Also, make
the assumption that if the required valve size is less than 6-inch
(DN 150), it will be installed using concentric reducers. Determine
the appropriate Design ED valve with a linear cage.
1. Specify the necessary variables required to size the valve:
a. Desired valve design—ASME CL300 Design ED valve
with a linear cage. Assume valve size is 4 inches.
b. Process uid—superheated steam
c. Service conditions—
w = 125 000 pounds/hr (56 700 kg/hr)
P1 = 500 psig (34,5 bar) = 514.7 psia (35,5 bar)
P2 = 250 psig (17 bar) = 264.7 psia (18,3 bar)
P = 250 psi (17 bar)
x = P/P1 = 250/514.7 = 0.49
T1 = 500°F (260°C)
γ1 = 1.0434 pound/ft3 (16,71 kg/m3)
(from Properties of Saturated Steam Table)
k = 1.28 (from Properties of Saturated Steam Table)
Cv =
q
N7 FP P1 Y
Gg T1 Z
x
Cv = = 1515
6.0 x 10
6
(1360)(1.0)(214.7)(0.667)
(0.6)(520)(1.0)
0.129
Cv =
q
N7 FP P1 Y
Gg T1 Z
x
=
6.0 x 10
6
(1360)(1.0)(214.7)(0.667)
(0.6)(520)(1.0)
0.237
= 1118
Cv =
q
N7 FP P1 Y
Gg T1 Z
x
=
6.0 x 10
6
(1360)(1.0)(214.7)(0.667)
(0.6)(520)(1.0)
0.308
= 980
Page 74
Valve Sizing (Standardized Method)
644
Te c h n i c a l
Because the 4-inch valve is to be installed in a 6-inch line, the xT
term must be replaced by xTP.
where,
N5 = 1000, from the Equation Constants Table
d = 4 inches
Fp = 0.95, determined in step 3
xT = 0.688, a value determined from the appropriate
listing in the ow coefcient table
Cv = 236, from step 3
and
Ki = K1 + K
B1
= 0.96
where D = 6-inch
so:
Finally:
5. Solve for required Cv using the appropriate equation.
2. Determine the appropriate equation constant, N, from the
Equation Constants Table 3-2 in Liquid Valve Sizing Section.
Because the specied ow rate is in mass units, (pound/hr), and
the specic weight of the steam is also specied, the only sizing
equation that can be used is that which contains the N6 constant.
Therefore,
N6 = 63.3
3. Determine Fp, the piping geometry factor.
where,
N2 = 890, determined from the Equation Constants Table
d = 4 inches
Cv = 236, which is the value listed in the ow coefcient
Table 4-3 for a 4-inch Design ED valve at 100%
total travel.
Σ
K = K1 + K
2
Finally,
4. Determine Y, the expansion factor.
where,
Fp = 1 +
Σ
K
N
2
Cv
d2
2
-1/2
= 1.5 1 -
d2
D2
2
= 1.5 1 -
42
62
2
= 0.463
Fp = 1 +
0.463
(1.0)(236)
2
-1/2
890
(4)
2
= 0.95
Y = 1 -
x
3Fkx
TP
Fk =
k
1.40
=
1.40
1.28
= 0.91
x = 0.49 (As calculated in step 1.)
xTP = 1 +
xT
F
p
2
xT K
i
N
5
Cv
d2
2
-1
= 0.5 1 - + 1 -
d
2
D2 D
4
d
= 0.5 1 - + 1 -
4
2
62
6
4
4
xTP = 1 + = 0.67
0.69
236
42
2
-1
0.952
(0.69)(0.96)
1000
Y = 1 -
x
3 F
k xTP
= 1 -
0.49
(3) (0.91) (0.67)
= 0.73
Cv =
w
N6 FP Y
x P1 γ
1
=
125,000
(63.3)(0.95)(0.73)
(0.49)(514.7)(1.0434)
= 176
2
2
Page 75
Valve Sizing (Standardized Method)
645
Te c h n i c a l
Table 4-2. Representative Sizing Coefcients for Rotary Shaft Valves
VALVE SIZE,
INCHES
VALVE STYLE
DEGREES OF
VALVE OPENING
C
V
F
L
X
T
F
D
1
V-Notch Ball Valve 60
90
15.6
34.0
0.86
0.86
0.53
0.42
- - - -
- - - -
1-1/2
V-Notch Ball Valve 60
90
28.5
77.3
0.85
0.74
0.50
0.27
- - - -
- - - -
2
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
59.2
132
58.9
80.2
0.81
0.77
0.76
0.71
0.53
0.41
0.50
0.44
- - - -
- - - -
0.49
0.70
3
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
120
321
115
237
0.80
0.74
0.81
0.64
0.50
0.30
0.46
0.28
0.92
0.99
0.49
0.70
4
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
195
596
270
499
0.80
0.62
0.69
0.53
0.52
0.22
0.32
0.19
0.92
0.99
0.49
0.70
6
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
340
1100
664
1260
0.80
0.58
0.66
0.55
0.52
0.20
0.33
0.20
0.91
0.99
0.49
0.70
8
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
518
1820
1160
2180
0.82
0.54
0.66
0.48
0.54
0.18
0.31
0.19
0.91
0.99
0.49
0.70
10
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
1000
3000
1670
3600
0.80
0.56
0.66
0.48
0.47
0.19
0.38
0.17
0.91
0.99
0.49
0.70
12
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
1530
3980
2500
5400
0.78
0.63
- - - -
- - - -
0.49
0.25
- - - -
- - - -
0.92
0.99
0.49
0.70
16
V-Notch Ball Valve
High Performance Buttery Valve
60
90
60
90
2380
8270
3870
8600
0.80
0.37
0.69
0.52
0.45
0.13
0.40
0.23
0.92
1.00
- - - -
- - - -
Table 4-1. Representative Sizing Coefcients for Type 1098-EGR Regulator
BODY SIZE,
INCHES (DN)
LINEAR CAGE
Line Size Equals Body Size 2:1 Line Size to Body Size
X
T
F
D
F
L
C
v
C
v
Regulating Wide-Open Regulating Wide-Open
1 (25) 16.8 17.7 17.2 18.1 0.806 0.43
0.84
2 (50) 63.3 66.7 59.6 62.8 0.820 0.35
3 (80) 132 139 128 135 0.779 0.30
4 (100) 202 213 198 209 0.829 0.28
6 (150) 397 418 381 404 0.668 0.28
BODY SIZE,
INCHES (DN)
WHISPER TRIMTM CAGE
Line Size Equals Body Size Piping 2:1 Line Size to Body Size Piping
X
T
F
D
F
L
C
v
C
v
Regulating Wide-Open Regulating Wide-Open
1 (25) 16.7 17.6 15.6 16.4 0.753 0.10
0.89
2 (50) 54 57 52 55 0.820 0.07
3 (80) 107 113 106 110 0.775 0.05
4 (100) 180 190 171 180 0.766 0.04
6 (150) 295 310 291 306 0.648 0.03
Page 76
Valve Sizing (Standardized Method)
646
Te c h n i c a l
6. Select the valve size using ow coefcient tables and the
calculated C
v
value.
Refer to the ow coefcient Table 4-3 for Design ED valves with
linear cage. Because the assumed 4-inch valve has a Cv of 236 at
100% travel and the next smaller size (3-inch) has a Cv of
only 148, it can be surmised that the assumed size is correct. In
the event that the calculated required Cv had been small enough
to have been handled by the next smaller size, or if it had been
larger than the rated Cv for the assumed size, it would have been
necessary to rework the problem again using values for the new
assumed size.
7. Sizing equations for compressible uids.
The equations listed below identify the relationships between
ow rates, ow coefcients, related installation factors, and
pertinent service conditions for control valves handling
compressible uids. Flow rates for compressible uids may
be encountered in either mass or volume units and thus equations
are necessary to handle both situations. Flow coefcients may be
calculated using the appropriate equations selected from the
following. A sizing ow chart for compressible uids is given in
Annex B.
The ow rate of a compressible uid varies as a function of the
ratio of the pressure differential to the absolute inlet pressure
( P/P1), designated by the symbol x. At values of x near zero,
the equations in this section can be traced to the basic Bernoulli
equation for Newtonian incompressible uids. However,
increasing values of x result in expansion and compressibility
effects that require the use of appropriate factors (see Buresh,
Schuder, and Driskell references).
7.1 Turbulent ow
7.1.1 Non-choked turbulent ow
7.1.1.1 Non-choked turbulent ow without attached ttings
[Applicable if x < FγxT]
The ow coefcient shall be calculated using one of the
following equations:
Eq. 6
Eq. 7
Eq. 8a
Eq. 8b
NOTE 1 Refer to 8.5 for details of the expansion factor Y.
NOTE 2 See Annex C for values of M.
7.1.1.2 Non-choked turbulent ow with attached ttings
[Applicable if x < FγxTP]
C =
W
N6Y
xP1ρ
1
C =
N8P1Y
W
xM
T1Z
C =
N9P1Y
Q
x
MT1Z
C =
N7P1Y
Q
x
GgT1Z
Table 4-3. Representative Sizing Coefcients for Design ED Single-Ported Globe Style Valve Bodies
VALVE SIZE,
INCHES
VALVE PLUG
STYLE
FLOW
CHARACTERISTICS
PORT DIAMETER,
INCHES (mm)
RATED TRAVEL,
INCHES (mm)
C
V
F
L
X
T
F
D
1/2 Post Guided Equal Percentage 0.38 (9,7) 0.50 (12,7) 2.41 0.90 0.54 0.61
3/4 Post Guided Equal Percentage 0.56 (14,2) 0.50 (12,7) 5.92 0.84 0.61 0.61
1
Micro-Form
TM
Cage Guided
Equal Percentage
Linear
Equal Percentage
3/8
1/2
3/4
1-5/16
1-5/16
(9,5)
(12,7)
(19,1)
(33,3)
(33,3)
3/4
3/4
3/4
3/4
3/4
(19,1)
(19,1)
(19,1)
(19,1)
(19,1)
3.07
4.91
8.84
20.6
17.2
0.89
0.93
0.97
0.84
0.88
0.66
0.80
0.92
0.64
0.67
0.72
0.67
0.62
0.34
0.38
1-1/2
Micro-Form
TM
Cage Guided
Equal Percentage
Linear
Equal Percentage
3/8
1/2
3/4
1-7/8
1-7/8
(9,5)
(12,7)
(19,1)
(47,6)
(47,6)
3/4
3/4
3/4
3/4
3/4
(19,1)
(19,1)
(19,1)
(19,1)
(19,1)
3.20
5.18
10.2
39.2
35.8
0.84
0.91
0.92
0.82
0.84
0.65
0.71
0.80
0.66
0.68
0.72
0.67
0.62
0.34
0.38
2 Cage Guided
Linear
Equal Percentage
2-5/16
2-5/16
(58,7)
(58,7)
1-1/8
1-1/8
(28,6)
(28,6)
72.9
59.7
0.77
0.85
0.64
0.69
0.33
0.31
3 Cage Guided
Linear
Equal Percentage
3-7/16 (87,3) 1-1/2 (38,1) 148
136
0.82
0.82
0.62
0.68
0.30
0.32
- - - - - - - -
4 Cage Guided
Linear
Equal Percentage
4-3/8 (111) 2 (50,8) 236
224
0.82
0.82
0.69
0.72
0.28
0.28
- - - - - - - -
6 Cage Guided
Linear
Equal Percentage
7 (178) 2 (50,8) 433
394
0.84
0.85
0.74
0.78
0.28
0.26
- - - - - - - -
8 Cage Guided
Linear
Equal Percentage
8 (203) 3 (76,2) 846
818
0.87
0.86
0.81
0.81
0.31
0.26
- - - - - - - -
Page 77
Cold Temperature Considerations
647
Te c h n i c a l
Regulators Rated for Low Temperatures
In some areas of the world, regulators periodically operate in
temperatures below -20°F (-29°C). These cold temperatures
require special construction materials to prevent regulator failure.
Emerson Process Management offers regulator constructions that
are RATED for use in service temperatures below -20°F (-29°C).
Selection Criteria
When selecting a regulator for extreme cold temperature service,
the following guidelines should be considered:
• The body material should be 300 Series stainless steel, LCC,
or LCB due to low carbon content in the material makeup.
• Give attention to the bolts used. Generally, special stainless
steel bolting is required.
• Gaskets and O-rings may need to be addressed if providing
a seal between two parts exposed to the cold.
• Special springs may be required in order to prevent fracture
when exposed to extreme cold.
• Soft parts in the regulator that are also being used as a seal
gasket between two metal parts (such as a diaphragm) may
need special consideration. Alternate diaphragm materials
should be used to prevent leakage caused by hardening and
stiffening of the standard materials.
Page 78
Freezing
648
Te c h n i c a l
Introduction
Freezing has been a problem since the birth of the gas industry.
This problem will likely continue, but there are ways to minimize
the effects of the phenomenon.
There are two areas of freezing. The rst is the formation of ice
from water travelling within the gas stream. Ice will form when
temperatures drop below 32°F (0°C).
The second is hydrate formation. Hydrate is a frozen mixture
of water and hydrocarbons. This bonding of water around the
hydrocarbon molecule forms a compound which can freeze above
32°F (0°C). Hydrates can be found in pipelines that are saturated
with water vapor. It is also common to have hydrate formation in
natural gas of high BTU content. Hydrate formation is dependent
upon operating conditions and gas composition.
Reducing Freezing Problems
To minimize problems, we have several options.
1. Keep the uid temperature above the freezing point by
applying heat.
2. Feed an antifreeze solution into the ow stream.
3. Select equipment that is designed to be ice-free in the
regions where there are moving parts.
4. Design systems that minimize freezing effects.
5. Remove the water from the ow stream.
Heat the Gas
Obviously, warm water does not freeze. What we need to know is
when is it necessary to provide additional heat.
Gas temperature is reduced whenever pressure is reduced. This
temperature drop is about 1°F (-17°C) for each 15 psi (1,03 bar)
pressure drop. Potential problems can be identied by calculating
the temperature drop and subtracting from the initial temperature.
Usually ground temperature, about 50°F (10°C) is the initial
temperature. If a pressure reducing station dropped the pressure
from 400 to 250 psi (28 to 17 bar) and the initial temperature is
50°F (10°C), the nal temperature would be 40°F (4°C).
50°F - (400 to 250 psi) (1°F/15 psi) = 40°F
(10°C - (28 to 17 bar) (-17°C/1,03 bar) = 5°C)
In this case, a freezing problem is not expected. However, if the
nal pressure was 25 psi (1,7 bar) instead of 250 psi (17 bar),
the nal temperature would be 25°F (-4°C). We should expect
freezing in this example if there is any moisture in the gas stream.
We can heat the entire gas stream with line heaters where the
situation warrants. However, this does involve some large
equipment and considerable fuel requirements.
Many different types of large heaters are on the market today. Some
involve boilers that heat a water/glycol solution which is circulated
through a heat exchanger in the main gas line. Two important
considerations are: (1) fuel efciencies, and (2) noise generation.
In many cases, it is more practical to build a box around the
pressure reducing regulator and install a small catalytic heater
to warm the regulator. When pilot-operated regulators are used,
we may nd that the ice passes through the regulator without
difculty but plugs the small ports in the pilot. A small heater can
be used to heat the pilot supply gas or the pilot itself. A word of
caution is appropriate. When a heater remains in use when it is not
needed, it can overheat the rubber parts of the regulator. They are
usually designed for 180°F (82°C) maximum. Using an automatic
temperature control thermostat can prevent overheating.
Antifreeze Solution
An antifreeze solution can be introduced into the ow stream where
it will combine with the water. The mixture can pass through
the pressure reducing station without freezing. The antifreeze is
dripped into the pipeline from a pressurized reservoir through a
needle valve. This system is quite effective if one remembers to
replenish the reservoir. There is a system that allows the antifreeze
to enter the pipeline only when needed. We can install a small
pressure regulator between the reservoir and the pipeline with the
control line of the small regulator connected downstream of the
pressure reducing regulator in the pipeline. The small regulator is
set at a lower pressure than the regulator in the pipeline. When the
controlled pressure is normal, the small regulator remains closed and
conserves the antifreeze. When ice begins to block the regulator in
the pipeline, downstream pressure will fall below the setpoint of the
small regulator which causes it to open, admitting antifreeze into the
pipeline as it is needed. When the ice is removed, the downstream
pressure returns to normal and the small regulator closes until ice
begins to re-form. This system is quite reliable as long as the supply
of the antifreeze solution is maintained. It is usually used at low
volume pressure reducing stations.
Equipment Selection
We can select equipment that is somewhat tolerant of freezing if
we know how ice forms in a pressure reducing regulator. Since
the pressure drop occurs at the orice, this is the spot where we
might expect the ice formation. However, this is not necessarily
the case. Metal regulator bodies are good heat conductors. As a
result, the body, not just the port, is cooled by the pressure drop.
The moisture in the incoming gas strikes the cooled surface as it
enters the body and freezes to the body wall before it reaches the
orice. If the valve plug is located upstream of the orice, there is
a good chance that it will become trapped in the ice and remain in
the last position. This ice often contains worm holes which allow
Page 79
Freezing
649
Te c h n i c a l
gas to continue to ow. In this case, the regulator will be unable to
control downstream pressure when the ow requirement changes.
If the valve plug is located downstream of the port, it is operating
in an area that is frequently ice-free. It must be recognized that
any regulator can be disabled by ice if there is sufcient moisture
in the ow stream.
System Design
We can arrange station piping to reduce freezing if we know
when to expect freezing. Many have noted that there are few
reported instances of freezing when the weather is very cold (0°F
(-18°C)). They have observed that most freezing occurs when the
atmospheric temperature is between 35° and 45°F (2° and 7°C).
When the atmospheric temperature is quite low, the moisture
within the gas stream freezes to the pipe wall before it reaches the
pressure reducing valve which leaves only dry gas to pass through
the valve. We can take advantage of this concept by increasing
the amount of piping that is exposed above ground upstream of the
pressure reducing valve. This will assure ample opportunity for
the moisture to contact the pipe wall and freeze to the wall.
When the atmospheric temperature rises enough to melt the ice
from the pipe wall, it is found that the operating conditions are not
favorable to ice formation in the pressure reducing valve. There
may be sufcient solar heat gain to warm the regulator body or
lower ow rates which reduces the refrigeration effect of the
pressure drop.
Parallel pressure reducing valves make a practical antifreeze system
for low ow stations such as farm taps. The two parallel regulators
are set at slightly different pressures (maybe one at 50 psi (3 bar)
and one at 60 psi (4 bar)). The ow will automatically go through
the regulator with the higher setpoint. When this regulator freezes
closed, the pressure will drop and the second regulator will open
and carry the load. Since most freezing instances occur when the
atmospheric temperature is between 35° and 45°F (2° and 7°C),
we expect the ice in the rst regulator to begin thawing as soon as
the ow stops. When the ice melts from the rst regulator, it will
resume owing gas. These two regulators will continue to alternate
between owing and freezing until the atmospheric temperature
decreases or increases, which will get the equipment out of the ice
formation temperature range.
Water Removal
Removing the moisture from the ow stream solves the problem
of freezing. However, this can be a difcult task. Where moisture
is a signicant problem, it may be benecial to use a method of
dehydration. Dehydration is a process that removes the water
from the gas stream. Effective dehydration removes enough water
to prevent reaching the dew point at the lowest temperature and
highest pressure.
Two common methods of dehydration involve glycol absorption
and desiccants. The glycol absorption process requires the gas
stream to pass through glycol inside a contactor. Water vapor
is absorbed by the glycol which in turn is passed through a
regenerator that removes the water by distillation. The glycol
is reused after being stripped of the water. The glycol system is
continuous and fairly low in cost. It is important, however, that
glycol is not pushed downstream with the dried gas.
The second method, solid absorption or desiccant, has the ability to
produce much drier gas than glycol absorption. The solid process
has the gas stream passing through a tower lled with desiccant.
The water vapor clings to the desiccant, until it reaches saturation.
Regeneration of the desiccant is done by passing hot gas through
the tower to dry the absorption medium. After cooling, the system
is ready to perform again. This is more of a batch process and will
require two or more towers to keep a continuous ow of dry gas.
The desiccant system is more expensive to install and operate than
the glycol units.
Most pipeline gas does not have water content high enough to
require these measures. Sometimes a desiccant dryer installed
in the pilot gas supply lines of a pilot-operated regulator is quite
effective. This is primarily true where water is present on an
occasional basis.
Summary
It is ideal to design a pressure reducing station that will never
freeze, but anyone who has spent time working on this problem
will acknowledge that no system is foolproof. We can design
systems that minimize the freezing potential by being aware of the
conditions that favor freezing.
Page 80
Sulfide Stress Cracking
--NACE MR0175-2002, MR0175/ISO 15156
650
Te c h n i c a l
The Details
NACE MR0175, “Sulde Stress Corrosion Cracking Resistant
Metallic Materials for Oil Field Equipment” is widely used
throughout the world. In late 2003, it became NACE MR0175/
ISO 15156, “Petroleum and Natural Gas Industries - Materials for
Use in H2S-Containing Environments in Oil and Gas Production.”
These standards specify the proper materials, heat treat conditions
and strength levels required to provide good service life in sour
gas and oil environments.
NACE International (formerly the National Association of
Corrosion Engineers) is a worldwide technical organization which
studies various aspects of corrosion and the damage that may
result in reneries, chemical plants, water systems and other types
of industrial equipment. MR0175 was rst issued in 1975, but the
origin of the document dates to 1959 when a group of engineers
in Western Canada pooled their experience in successful handling
of sour gas. The group organized as a NACE committee and in
1963 issued specication 1B163, “Recommendations of Materials
for Sour Service.” In 1965, NACE organized a nationwide
committee, which issued 1F166 in 1966 and MR0175 in 1975.
Revisions were issued on an annual basis as new materials and
processes were added. Revisions had to receive unanimous
approval from the responsible NACE committee.
In the mid-1990’s, the European Federation of Corrosion (EFC)
issued 2 reports closely related to MR0175; Publication 16,
“Guidelines on Materials Requirements for Carbon and Low Alloy
Steels for H2S-Containing Environments in Oil and Gas Production”
and Publication 17, “Corrosion Resistant Alloys for Oil and Gas
Production: Guidance on General Requirements and Test Methods
for H2S Service.” EFC is located in London, England.
The International Organization for Standardization (ISO) is a
worldwide federation of national standards bodies from more than
140 countries. One organization from each country acts as the
representative for all organizations in that country. The American
National Standards Institute (ANSI) is the USA representative
in ISO. Technical Committee 67, “Materials, Equipment and
Offshore Structures for Petroleum, Petrochemical and Natural Gas
Industries,” requested that NACE blend the different sour service
documents into a single global standard.
This task was completed in late 2003 and the document was
issued as ISO standard, NACE MR0175/ISO 15156. It is
now maintained by ISO/TC 67, Work Group 7, a 12-member
“Maintenance Panel” and a 40-member Oversight Committee
under combined NACE/ISO control. The three committees are an
international group of users, manufacturers and service providers.
Membership is approved by NACE and ISO based on technical
knowledge and experience. Terms are limited. Previously, some
members on the NACE Task Group had served for over 25 years.
NACE MR0175/ISO 15156 is published in 3 volumes.
Part 1: General Principles for Selection of Cracking-Resistant Materials
Part 2: Cracking-Resistant Carbon and Low Alloy Steels, and the
Use of Cast Irons
Part 3: Cracking-Resistant CRA’s (Corrosion-Resistant Alloys) and
Other Alloys
NACE MR0175/ISO 15156 applies only to petroleum production,
drilling, gathering and ow line equipment and eld processing
facilities to be used in H2S bearing hydrocarbon service. In the
past, MR0175 only addressed sulde stress cracking (SSC). In
NACE MR0175/ISO 15156, however, but both SSC and chloride
stress corrosion cracking (SCC) are considered. While clearly
intended to be used only for oil eld equipment, industry has
applied MR0175 in to many other areas including reneries, LNG
plants, pipelines and natural gas systems. The judicious use of
the document in these applications is constructive and can help
prevent SSC failures wherever H2S is present. Saltwater wells and
saltwater handling facilities are not covered by NACE MR0175/
ISO 15156. These are covered by NACE Standard RP0475,
“Selection of Metallic Materials to Be Used in All Phases of Water
Handling for Injection into Oil-Bearing Formations.”
When new restrictions are placed on materials in NACE MR0175/
ISO 15156 or when materials are deleted from this standard,
materials in use at that time are in compliance. This includes
materials listed in MR0175-2002, but not listed in NACE
MR0175/ISO 15156. However, if this equipment is moved to a
different location and exposed to different conditions, the materials
must be listed in the current revision. Alternatively, successful use
of materials outside the limitations of NACE MR0175/ISO 15156
may be perpetuated by qualication testing per the standard. The
user may replace materials in kind for existing wells or for new
wells within a given eld if the environmental conditions of the
eld have not changed.
Page 81
Sulfide Stress Cracking
--NACE MR0175-2002, MR0175/ISO 15156
651
Te c h n i c a l
New Sulfide Stress Cracking Standard
for Refineries
Don Bush, Principal Engineer - Materials, at Emerson Process
Management Fisher Valves, is a member and former chair of
a NACE task group that has written a document for renery
applications, NACE MR0103. The title is “Materials Resistant
to Sulde Stress Cracking in Corrosive Petroleum Rening
Environments.” The requirements of this standard are very similar
to the pre-2003 MR0175 for many materials. When applying this
standard, there are changes to certain key materials compared with
NACE MR0175-2002.
Responsibility
It has always been the responsibility of the end user to determine
the operating conditions and to specify when NACE MR0175
applies. This is now emphasized more strongly than ever in NACE
MR0175/ISO 15156. The manufacturer is responsible for meeting
the metallurgical requirements of NACE MR0175/ISO 15156. It
is the end user’s responsibility to ensure that a material will be
satisfactory in the intended environment. Some of the operating
conditions which must be considered include pressure, temperature,
corrosiveness, uid properties, etc. When bolting components
are selected, the pressure rating of anges could be affected. It
is always the responsibility of the equipment user to convey the
environmental conditions to the equipment supplier, particularly if
the equipment will be used in sour service.
The various sections of NACE MR0175/ISO 15156 cover the
commonly available forms of materials and alloy systems. The
requirements for heat treatment, hardness levels, conditions of
mechanical work and post-weld heat treatment are addressed for
each form of material. Fabrication techniques, bolting, platings and
coatings are also addressed.
Applicability of NACE MR0175/ISO 15156
Low concentrations of H2S (<0.05 psi (0,3 kPa) H2S partial
pressure) and low pressures (<65 psia or 450 kPa) are considered
outside the scope of NACE MR0175/ISO 15156. The low stress
levels at low pressures or the inhibitive effects of oil may give
satisfactory performance with standard commercial equipment.
Many users, however, have elected to take a conservative approach
and specify compliance to either NACE MR0175 or NACE
MR0175/ISO 15156 any time a measurable amount of H2S is
present. The decision to follow these specications must be
made by the user based on economic impact, the safety aspects
should a failure occur and past eld experience. Legislation can
impact the decision as well. Such jurisdictions include; the Texas
Railroad Commission and the U.S. Minerals Management Service
(offshore). The Alberta, Canada Energy Conservation Board
recommends use of the specications.
Basics of Sulfide Stress Cracking (SSC) and Stress
Corrosion Cracking (SCC)
SSC and SCC are cracking processes that develop in the presence
of water, corrosion and surface tensile stress. It is a progressive
type of failure that produces cracking at stress levels that are
well below the material’s tensile strength. The break or fracture
appears brittle, with no localized yielding, plastic deformation or
elongation. Rather than a single crack, a network of ne, feathery,
branched cracks will form (see Figure 1). Pitting is frequently
seen, and will serve as a stress concentrator to initiate cracking.
With SSC, hydrogen ions are a product of the corrosion process
(Figure 2). These ions pick up electrons from the base material
producing hydrogen atoms. At that point, two hydrogen atoms
may combine to form a hydrogen molecule. Most molecules
will eventually collect, form hydrogen bubbles and oat away
harmlessly. However, some percentage of the hydrogen atoms will
diffuse into the base metal and embrittle the crystalline structure.
When a certain critical concentration of hydrogen is reached and
combined with a tensile stress exceeding a threshold level, SSC
will occur. H2S does not actively participate in the SSC reaction;
however, suldes act to promote the entry of the hydrogen atoms
into the base material.
As little as 0.05 psi (0,3 kPa) H2S partial pressure in 65 psia
(450 kPa) hydrocarbon gas can cause SSC of carbon and low
alloy steels. Sulde stress cracking is most severe at ambient
temperature, particularly in the range of 20° to 120°F (-6° to
49°C). Below 20°F (-6°C) the diffusion rate of the hydrogen is
Figure 1. Photomicrograph Showing Stress Corrosion Cracking
Page 82
Sulfide Stress Cracking
--NACE MR0175-2002, MR0175/ISO 15156
652
Te c h n i c a l
so slow that the critical concentration is never reached. Above
120°F (49°C), the diffusion rate is so fast that the hydrogen
atoms pass through the material in such a rapid manner that the
critical concentration is not reached.
Chloride SCC is widely encountered and has been extensively
studied. Much is still unknown, however, about its mechanism.
One theory says that hydrogen, generated by the corrosion
process, diffuses into the base metal in the atomic form and
embrittles the lattice structure. A second, more widely accepted
theory proposes an electrochemical mechanism. Stainless steels
are covered with a protective, chromium oxide lm. The chloride
ions rupture the lm at weak spots, resulting in anodic (bare)
and cathodic (lm covered) sites. The galvanic cell produces
accelerated attack at the anodic sites, which when combined with
tensile stresses produces cracking. A minimum ion concentration
is required to produce SCC. As the concentration increases, the
environment becomes more severe, reducing the time to failure.
Temperature also is a factor in SCC. In general, the likelihood of
SCC increases with increasing temperature. A minimum threshold
temperature exists for most systems, below which SCC is rare.
Across industry, the generally accepted minimum temperature for
chloride SCC of the 300 SST’s is about 160°F (71°C). NACE
MR0175/ISO 15156 has set a very conservative limit of 140°F
(60°C) due to the synergistic effects of the chlorides, H2S and low
pH values. As the temperature increases above these values, the
time to failure will typically decrease.
Resistance to chloride SCC increases with higher alloy materials.
This is reected in the environmental limits set by NACE
MR0175/ISO 15156. Environmental limits progressively increase
from 400 Series SST and ferritic SST to 300 Series, highly alloy
austenitic SST, duplex SST, nickel and cobalt base alloys.
Carbon Steel
Carbon and low-alloy steels have acceptable resistance to SSC and
SCC however; their application is often limited by their low resistance
to general corrosion. The processing of carbon and low alloy steels
must be carefully controlled for good resistance to SSC and SCC. The
hardness must be less than 22 HRC. If welding or signicant cold
working is done, stress relief is required. Although the base metal
hardness of a carbon or alloy steel is less than 22 HRC, areas of the
heat affected zone (HAZ) will be harder. PWHT will eliminate these
excessively hard areas.
ASME SA216 Grades WCB and WCC and SAME SA105 are
the most commonly used body materials. It is Fisher’s policy to
stress relieve all welded carbon steels that are supplied to NACE
MR0175/ISO 15156.
All carbon steel castings sold to NACE MR0175/ISO 15156
requirements are produced using one of the following processes:
1. In particular product lines where a large percentage of
carbon steel assemblies are sold as NACE MR0175/ISO
15156 compliant, castings are ordered from the foundry with
a requirement that the castings be either normalized or stress
relieved following all weld repairs, major or minor. Any
weld repairs performed, either major or minor, are
subsequently stress relieved.
2. In product lines where only a small percentage of carbon steel
products are ordered NACE MR0175/ISO 15156 compliant,
stock castings are stress relieved whether they are weld repaired
by Emerson Process Management or not. This eliminates the
chance of a minor foundry weld repair going undetected and not
being stress relieved.
ASME SA352 grades LCB and LCC have the same composition
as WCB and WCC, respectively. They are heat treated differently
and impact tested at -50°F (-46°C) to ensure good toughness in low
temperature service. LCB and LCC are used in locations where
temperatures commonly drop below the -20°F (-29°C) permitted for
WCB and WCC. LCB and LCC castings are processed in the same
manner as WCB and WCC when required to meet NACE MR0175/
ISO 15156.
For carbon and low-alloy steels NACE MR0175/ISO 15156
imposes some changes in the requirements for the weld procedure
qualication report (PQR). All new PQR’s will meet these
requirements; however, it will take several years for Emerson
Process Management and our suppliers to complete this work. At
this time, we will require user approval to use HRC.
H
H
H
H
H
H
H
H
H
H H
H
H
H
+
H
+
M
+
S
-2
H
+
H
+
H
+
H
+
S
-2
H
H
H
2
M
+
S
-2
M
+
e
-
e
-
e
-
σ
σ
Figure 2. Schematic Showing the Generation of Hydrogen Producing SSC
Page 83
Sulfide Stress Cracking
--NACE MR0175-2002, MR0175/ISO 15156
653
Te c h n i c a l
Carbon and Low-Alloy Steel Welding
Hardness Requirements
• HV-10, HV-5 or Rockwell 15N.
• HRC testing is acceptable if the design stresses are less than
67% of the minimum specied yield strength and the PQR
includes PWHT.
• Other methods require user approval.
• 250 HV or 70.6 HR15N maximum.
• 22 HRC maximum if approved by user.
Low-Alloy Steel Welding Hardness Requirements
• All of the above apply with the additional requirement of stress
relieve at 1150°F (621°C) minimum after welding.
All new PQR’s at Emerson Process Management and our foundries
will require hardness testing with HV-10, HV-5 or Rockwell 15N
and HRC. The acceptable maximum hardness values will be 250
HV or 70.6 HR15N and 22 HRC. Hardness traverse locations are
specied in NACE MR0175/ISO 15156 part 2 as a function of
thickness and weld conguration. The number and locations of
production hardness tests are still outside the scope of the standard.
The maximum allowable nickel content for carbon and low-alloy
steels and their weld deposits is 1%.
Low alloy steels like WC6, WC9, and C5 are acceptable to NACE
MR0175/ISO 15156 to a maximum hardness of 22 HRC. These
castings must all be stress relieved to FMS 20B52.
The compositions of C12, C12a, F9 and F91 materials do not fall
within the denition of “low alloy steel” in NACE MR0175/ISO
15156, therefore, these materials are not acceptable.
A few customers have specied a maximum carbon equivalent
(CE) for carbon steel. The primary driver for this requirement
is to improve the SSC resistance in the as-welded condition.
Fisher’s practice of stress relieving all carbon steel negates this
need. Decreasing the CE reduces the hardenability of the steel and
presumably improves resistance to sulde stress cracking (SSC).
Because reducing the CE decreases the strength of the steel, there
is a limit to how far the CE can be reduced.
Cast Iron
Gray, austenitic and white cast irons cannot be used for any
pressure-retaining parts, due to low ductility. Ferritic ductile iron
to ASTM A395 is acceptable when permitted by ANSI, API or
other industry standards.
Stainless Steel
400 Series Stainless Steel
UNS 410 (410 SST), CA15 (cast 410), 420 (420 SST) and several
other martensitic grades must be double tempered to a maximum
hardness of 22 HRC. PWHT is also required. An environmental
limit now applies to the martensitic grades; 1.5 psi (10 kPa) H2S
partial pressure and pH greater than or equal to 3.5, 416 (416 SST)
is similar to 410 (410) with the exception of a sulfur addition to
produce free machining characteristics. Use of 416 and other free
machining steels is not permitted by NACE MR0175/ISO 15156.
CA6NM is a modied version of the cast 410 stainless steel.
NACE MR0175/ISO 15156 allows its use, but species the
exact heat treatment required. Generally, the carbon content
must be restricted to 0.03% maximum to meet the 23 HRC
maximum hardness. PWHT is required for CA6NM. The same
environmental limit applies; 1.5 psi (10 kPa) H2S partial pressure
and pH greater than or equal to 3.5.
300 Series Stainless Steel
Several changes have been made with the requirements of the
austenitic (300 Series) stainless steels. Individual alloys are no
longer listed. All alloys with the following elemental ranges
are acceptable: C 0.08% maximum, Cr 16% minimum, Ni 8%
minimum, P 0.045% maximum, S 0.04% maximum, Mn 2.0%
maximum, and Si 2.0% maximum. Other alloying elements are
permitted. The other requirements remain; solution heat treated
condition, 22 HRC maximum and free of cold work designed to
improve mechanical properties. The cast and wrought equivalents of
302, 304 (CF8), S30403 (CF3), 310 (CK20), 316 (CF8M), S31603
(CF3M), 317 (CG8M), S31703 (CG3M), 321, 347 (CF8C) and
N08020 (CN7M) are all acceptable per NACE MR0175/ISO 15156.
Environmental restrictions now apply to the 300 Series SST. The
limits are 15 psia (100 kPa) H2S partial pressure, a maximum
temperature of 140°F (60°C), and no elemental sulfur. If the
chloride content is less than 50 mg/L (50 ppm), the H2S partial
pressure must be less than 50 psia (350 kPa) but there is no
temperature limit.
There is less of a restriction on 300 Series SST in oil and gas
processing and injection facilities. If the chloride content in
aqueous solutions is low (typically less than 50 mg/L or 50 ppm
chloride) in operations after separation, there are no limits for
austenitic stainless steels, highly alloyed austenitic stainless steels,
duplex stainless steels, or nickel-based alloys.
Page 84
Sulfide Stress Cracking
--NACE MR0175-2002, MR0175/ISO 15156
654
Te c h n i c a l
Post-weld heat treatment of the 300 Series SST is not required.
Although the corrosion resistance may be affected by poorly
controlled welding, this can be minimized by using the low carbon
ller material grades, low heat input levels and low interpass
temperatures. We impose all these controls as standard practice.
NACE MR0175/ISO 15156 now requires the use of “L” grade
consumables with 0.03% carbon maximum.
S20910
S20910 (Nitronic® 50) is acceptable in both the annealed and high
strength conditions with environmental restrictions; H2S partial
pressure limit of 15 psia (100 kPa), a maximum temperature
of 150°F (66°C), and no elemental sulfur. This would apply
to components such as bolting, plugs, cages, seat rings and
other internal parts. Strain hardened (cold-worked) S20910 is
acceptable for shafts, stems, and pins without any environmental
restrictions. Because of the environmental restrictions and poor
availability on the high strength condition, use of S20910 will
eventually be discontinued except for shafts, stems and pins where
unrestricted application is acceptable for these components.
CK3MCuN
The cast equivalent of S31254 (Avesta 254SMO®), CK3MCuN
(UNS J93254), is included in this category. The same elemental
limits apply. It is acceptable in the cast, solution heat-treated
condition at a hardness level of 100 HRB maximum in the absence
of elemental sulfur.
S17400
The use of S17400 (17-4PH) is now prohibited for pressureretaining components including bolting, shafts and stems. Prior to
2003, S17400 was listed as an acceptable material in the general
section (Section 3) of NACE MR0175. Starting with the 2003
revision, however, it is no longer listed in the general section. Its
use is restricted to internal, non-pressure containing components
in valves, pressure regulators and level controllers. This includes
cages and other trim parts. 17-4 bolting will no longer be supplied
in any NACE MR0175/ISO 15156 construction. The 17-4 and
15-5 must be heat-treated to the H1150 DBL condition or the
H1150M condition. The maximum hardness of 33 HRC is the
same for both conditions.
CB7Cu-1 and CB7Cu-2 (cast 17-4PH and 15-5 respectively) in
the H1150 DBL condition are also acceptable for internal valve
and regulator components. The maximum hardness is 30 HRC or
310 HB for both alloys.
Duplex Stainless Steel
Wrought and cast duplex SST alloys with 35-65% ferrite are
acceptable based on the composition of the alloy, but there are
environmental restrictions. There is no differentiation between cast
and wrought, therefore, cast CD3MN is now acceptable. There
are two categories of duplex SST. The “standard” alloys with a
30≤PREN≤40 and ≥1.5% Mo, and the “super” duplex alloys with
PREN>40. The PREN is calculated from the composition of the
material. The chromium, molybdenum, tungsten and nitrogen
contents are used in the calculation. NACE MR0175/ISO 15156
uses this number for several classes of materials.
PREN = Cr% + 3.3(Mo% + 0.5W%) + 16N%
The “standard” duplex SST grades have environmental limits of
450°F (232°C) maximum and H2S partial pressure of 1.5 psia (10 kPa)
maximum. The acceptable alloys include S31803, CD3MN, S32550
and CD7MCuN (Ferralium® 255). The alloys must be in the solution
heat-treated and quenched condition. There are no hardness restrictions
in NACE MR0175/ISO 15156, however, 28 HRC remains as the limit
in the renery document MR0103.
The “super” duplex SST with PREN>40 have environmental limits of
450°F (232°C) maximum and H2S partial pressure of 3 psia (20 kPa)
maximum. The acceptable “super” duplex SST’s include S32760 and
CD3MWCuN (Zeron® 100).
The cast duplex SST Z 6CNDU20.08M to the French National
Standard NF A 320-55 is no longer acceptable for NACE MR0175/
ISO 15156 applications. The composition fails to meet the
requirements set for either the duplex SST or the austenitic SST.
Highly Alloyed Austenitic Stainless Steels
There are two categories of highly alloyed austenitic SST’s that are
acceptable in the solution heat-treated condition. There are different
compositional and environmental requirements for the two categories.
The rst category includes alloys S31254 (Avesta 254SMO®) and
N08904 (904L); Ni% + 2Mo%>30 and Mo=2% minimum.
Alloy S31254 and N08904 Environmetal Limits
MAXIMUM
TEMPERATURE
MAXIMUM H2S
PARTIAL
PRESSURE
MAXIMUM
CHLORIDES
ELEMENTAL
SULFUR
140°F (60°C) 1.5 psia (10 kPa) No restriction No
140°F (60°C) 50 psia (345 kPa) 50 mg/L Chloride No
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Sulfide Stress Cracking
--NACE MR0175-2002, MR0175/ISO 15156
655
Te c h n i c a l
Alloy N07718 and N09925 Environmental Limits
MAXIMUM
TEMPERATURE
MAXIMUM H
L
2
S
PARTIAL PRESSURE
ELEMENTAL SULFUR
450°F (232°C) 30 psia (0,2 MPa) No
400°F (204°C) 200 psia (1,4 MPa) No
390°F (199°C) 330 psia (2,3 MPa) No
375°F (191°C) 360 psia (2,5 MPa) No
300°F (149°C) 400 psia (2,8 MPa) No
275°F (135°C) No limit Yes
Cast N07718 Environmental Limits
MAXIMUM
TEMPERATURE
MAXIMUM H
2
2
S
PARTIAL PRESSURE
ELEMENTAL
SULFUR
450°F (232°C) 30 psia (0,2 MPa) No
400°F (204°C) 200 psia (1,4 MPa) No
300°F (149°C) 400 psia (2,8 MPa) No
275°F (135°C) No limit Yes
Monel® K500 and Inconel® X750
N05500 and N07750 are now prohibited for use in pressureretaining components including bolting, shafts and stems. They
can still be used for internal parts such as cages, other trim
parts and torque tubes. There are no environmental restrictions,
however, for either alloy. They must be in the solution heat-treated
condition with a maximum hardness of 35 HRC. N07750 is still
acceptable for springs to 50 HRC maximum.
Cobalt-Base Alloys
Alloy 6 castings and hardfacing are still acceptable. There are
no environmental limits with respect to partial pressures of H2S
or elemental sulfur. All other cobalt-chromium-tungsten, nickelchromium-boron (Colmonoy) and tungsten-carbide castings are
also acceptable without restrictions.
The second category of highly alloyed austenitic stainless steels
are those having a PREN >40. This includes S31654 (Avesta
654SMO®), N08926 (Inco 25-6Mo), N08367 (AL-6XN), S31266
(UR B66) and S34565. The environmental restrictions for these
alloys are as follows:
Alloy S31654, N08926, N08367, S31266, and S34565
Environmental Limits
MAXIMUM
TEMPERATURE
MAXIMUM H
2
2
S
PARTIAL PRESSURE
MAXIMUM
CHLORIDES
ELEMENTAL
SULFUR
250°F (121°C) 100 psia (700 kPa) 5,000 mg/L chloride No
300°F (149°C) 45 psia (310 kPa) 5,000 mg/L chloride No
340°F (171°C) 15 psia (100 kPa) 5,000 mg/L chloride No
Nonferrous Alloys
Nickel-Base Alloys
Nickel base alloys have very good resistance to cracking in sour,
chloride containing environments. There are 2 different categories
of nickel base alloys in NACE MR0175/ISO 15156:
• Solid-solution nickel-based alloys
• Precipitation hardenable alloys
The solid solution alloys are the Hastelloy® C, Inconel® 625 and
Incoloy® 825 type alloys. Both the wrought and cast alloys are
acceptable in the solution heat-treated condition with no hardness
limits or environmental restrictions. The chemical composition of
these alloys is as follows:
• 19.0% Cr minimum, 29.5% Ni minimum, and 2.5% Mo
minimum. Includes N06625, CW6MC, N08825, CU5MCuC.
• 14.5% Cr minimum, 52% Ni minimum, and 12% Mo
minimum. Includes N10276, N06022, CW2M.
N08020 and CN7M (alloy 20 Cb3) are not included in this
category. They must follow the restrictions placed on the
austenitic SST’s like 304, 316 and 317.
Although originally excluded from NACE MR0175/ISO 15156,
N04400 (Monel® 400) in the wrought and cast forms are now
included in this category.
The precipitation hardenable alloys are Incoloy® 925, Inconel® 718
and X750 type alloys. They are listed in the specication as individual
alloys. Each has specic hardness and environmental restrictions.
N07718 is acceptable in the solution heat-treated and precipitation
hardened condition to 40 HRC maximum. N09925 is acceptable
in the cold-worked condition to 35 HRC maximum, solutionannealed and aged to 38 HRC maximum and cold-worked and
aged to 40 HRC maximum.
The restrictions are as follows:
Cast N07718 is acceptable in the solution heat-treated and
precipitation hardened condition to 35 HRC maximum. The
restrictions are as follows:
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656
Te c h n i c a l
All cobalt based, nickel based and tungsten-carbide weld overlays
are acceptable without environmental restrictions. This includes
CoCr-A, NiCr-A (Colmonoy® 4), NiCr-C (Colmonoy® 6) and
Haynes Ultimet® hardfacing.
Wrought UNS R31233 (Haynes Ultimet®) is acceptable in the solution
heat-treated condition to 22 HRC maximum, however, all production
barstock exceeds this hardness limit. Therefore, Ultimet® barstock
cannot be used for NACE MR0175/ISO 15156 applications. Cast
Ultimet is not listed in NACE MR0175/ISO 15156.
R30003 (Elgiloy®) springs are acceptable to 60 HRC in the cold
worked and aged condition. There are no environmental restrictions.
Aluminum and Copper Alloys
Per NACE MR0175/ISO 15156, environmental limits have not
been established for aluminum base and copper alloys. This
means that they could be used in sour applications per the
requirements of NACE MR0175/ISO 15156, however, they should
not be used because severe corrosion attack will likely occur.
They are seldom used in direct contact with H2S.
Titanium
Environmental limits have not been established for the wrought
titanium grades. Fisher® has no experience in using titanium in
sour applications. The only common industrial alloy is wrought
R50400 (grade 2). Cast titanium is not included in NACE
MR0175/ISO 15156.
Zirconium
Zirconium is not listed in NACE MR0175/ISO 15156.
Springs
Springs in compliance with NACE represent a difcult problem.
To function properly, springs must have very high strength
(hardness) levels. Normal steel and stainless steel springs would
be very susceptible to SSC and fail to meet NACE MR0175/ISO
15156. In general, relatively soft, low strength materials must
be used. Of course, these materials produce poor springs. The
two exceptions allowed are the cobalt based alloys, such as
R30003 (Elgiloy®), which may be cold worked and hardened to
a maximum hardness of 60 HRC and alloy N07750 (alloy X750)
which is permitted to 50 HRC. There are no environmental
restrictions for these alloys.
Coatings
Coatings, platings and overlays may be used provided the base
metal is in a condition which is acceptable per NACE MR0175/ISO
15156. The coatings may not be used to protect a base material
which is susceptible to SSC. Coatings commonly used in sour
service are chromium plating, electroless nickel (ENC) and
nitriding. Overlays and castings commonly used include CoCrA (Stellite® or alloy 6), R30006 (alloy 6B), NiCr-A and NiCr-C
(Colmonoy® 4 and 6) nickel-chromium-boron alloys. Tungsten
carbide alloys are acceptable in the cast, cemented or thermally
sprayed conditions. Ceramic coatings such as plasma sprayed
chromium oxide are also acceptable. As is true with all materials in
NACE MR0175/ISO 15156, the general corrosion resistance in the
intended application must always be considered.
NACE MR0175/ISO 15156 permits the uses of weld overlay
cladding to protect an unacceptable base material from cracking.
Fisher does not recommend this practice, however, as hydrogen
could diffuse through the cladding and produce cracking of a
susceptible basemetal such as carbon or low alloy steel.
Stress Relieving
Many people have the misunderstanding that stress relieving
following machining is required by NACE MR0175/ISO 15156.
Provided good machining practices are followed using sharp
tools and proper lubrication, the amount of cold work produced is
negligible. SSC and SCC resistance will not be affected. NACE
MR0175/ISO 15156 actually permits the cold rolling of threads,
provided the component will meet the heat treat conditions and
hardness requirements specied for the given parent material. Cold
deformation processes such as burnishing are also acceptable.
Bolting
Bolting materials must meet the requirements of NACE MR0175/
ISO 15156 when directly exposed to the process environment
(“exposed” applications). Standard ASTM A193 and ASME
SA193 grade B7 bolts or ASTM A194 and ASME SA194 grade
2H nuts can and should be used provided they are outside of
the process environment (“non-exposed” applications). If the
bolting will be deprived atmospheric contact by burial, insulation
or ange protectors and the customer species that the bolting
will be “exposed”, then grades of bolting such as B7 and 2H
are unacceptable. The most commonly used fasteners listed for
“exposed” applications are grade B7M bolts (99 HRB maximum)
and grade 2HM nuts (22 HRC maximum). If 300 Series SST
fasteners are needed, the bolting grades B8A Class 1A and B8MA
Class 1A are acceptable. The corresponding nut grades are 8A
and 8MA.
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--NACE MR0175-2002, MR0175/ISO 15156
657
Te c h n i c a l
It must be remembered, however, that the use of lower strength
bolting materials such as B7M may require pressure vessel
derating. The special S17400 double H1150 bolting previously
offered on E body valves to maintain the full B7 rating is no longer
acceptable to NACE MR0175/ISO 15156. Prior to the 2003,
S17400 was listed as an acceptable material in the general section
(Section 3) of NACE MR0175. Following the 2003 revision, it is
no longer listed in the general section. Its use is now restricted to
internal, non-pressure containing components in valves, pressure
regulators and level controllers. The use of S17400 for bolting is
specically prohibited. N07718 (alloy 718) bolting with 2HM nuts
is one alternative.
Two different types of packing box studs and nuts are commonly
used by Fisher®. The stainless steel type is B8M S31600 class 2
(strain hardened) and 316 nuts per FMS 20B86. The steel type is B7
studs with 2H nuts. If the customer species that the packing box
studs and nuts are “exposed” then grade B7M studs and grade 2HM
nuts or B8MA Class 1A studs and 8MA nuts are commonly used.
Bolting Coatings
NFC (Non-Corroding Finish) and ENC (Electroless Nickel
Coating) coatings are acceptable on pressure-retaining and nonpressure-retaining fasteners. For some reason, there is often
confusion regarding the acceptability of zinc plated fasteners per
NACE MR0175/ISO 15156. NACE MR0175/ISO 15156 does
not preclude the use of any coating, provided it is not used in an
attempt to prevent SSC or SCC of an otherwise unacceptable base
material. However, zinc plating of pressure-retaining bolting is not
recommended due to liquid metal induced embrittlement concerns.
Composition Materials
NACE MR0175/ISO 15156 does not address elastomer and
polymer materials although ISO/TC 67, Work Group 7 is now
working on a Part 4 to address these materials. The importance
of these materials in critical sealing functions, however, cannot be
overlooked. User experience has been successful with elastomers
such as Nitrile (NBR), Neoprene and the Fluoroelastomers (FKM)
and Peruoroelastomers (FFKM). In general, uoropolymers such
as Polytetrauoroethylene (PTFE), TCM Plus, TCM Ultra and
TCM III can be applied without reservation within their normal
temperature range.
Elastomer use is as follows:
1. If possible, use HNBR for sour natural gas, oil, or water at
temperatures below 250°F (121°C). It covers the widest
range of sour applications at a lower cost than PTFE or
Fluoroelastomer (FKM). Unfortunately, the material is
relatively new, and only a handful of parts are currently set
up. Check availability before specifying.
2. Use PTFE for sour natural gas, oil, or water applications
at temperatures between 250°F (121°C) and 400°F (204°C).
3. Fluoroelastomer (FKM) can be used for sour natural gas,
oil, or water applications with less than 10% H2S and
temperatures below 250°F (121°C).
4. Conventional Nitrile (NBR) can be used for sour natural
gas, oil, or water applications with less than 1% H2S and
temperatures below 150°F (66°C).
5. CR can be used for sour natural gas or water applications
involving temperatures below 150°F (66°C). Its resistance to
oil is not as good.
6. IIR and Ethylenepropylene (EPDM) (or EPR) can be used for
H2S applications that don’t involve hydrocarbons (H2S gas, sour
water, etc.).
Tubulars
A separate section has been established for downhole tubulars and
couplings. This section contains provisions for using materials in
the cold-drawn condition to higher hardness levels (cold-worked
to 35 HRC maximum). In some cases, the environmental limits
are also different. This has no affect on Fisher as we do not make
products for these applications. Nickel-based components used for
downhole casing, tubing, and the related equipment (hangers and
downhole component bodies; components that are internal to the
downhole component bodies) are subject to the requirements.
Expanded Limits and Materials
With documented laboratory testing and/or eld experience, it is
possible to expand the environmental limits of materials in NACE
MR0175/ISO 15156 or use materials not listed in NACE MR0175/
ISO 15156. This includes increasing the H2S partial pressure
limit or temperature limitations. Supporting documentation must
be submitted to NACE International Headquarters, which will
make the data available to the public. NACE International will
neither review nor approve this documentation. It is the user’s
responsibility to evaluate and determine the applicability of the
documented data for the intended application.
It is the user’s responsibility to ensure that the testing cited is
relevant for the intended applications. Choice of appropriate
temperatures and environments for evaluating susceptibility to
both SCC and SSC is required. NACE Standard TM0177 and EFC
Publication #1739 provide guidelines for laboratory testing.
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--NACE MR0175-2002, MR0175/ISO 15156
658
Te c h n i c a l
Field-based documentation for expanded alloy use requires
exposure of a component for sufcient time to demonstrate its
resistance to SCC/SSC. Sufcient information on factors that affect
SCC/SSC (e.g., stress levels, uid and gas composition, operating
conditions, galvanic coupling, etc.) must be documented.
Codes and Standards
Applicable ASTM, ANSI, ASME and API standards are used
along with NACE MR0175/ISO 15156 as they would normally
be used for other applications. The NACE MR0175/ISO 15156
requires that all weld procedures be qualied to these same
standards. Welders must be familiar with the procedures and
capable of making welds which comply.
Certification
Fisher® Certication Form 7508 is worded as follows for NACE
MR0175-2002 and MR0175/ISO 15156:
“NACE MR0175/ISO 15156 OR NACE MR0175-2002: This unit
meets the metallurgical requirements of NACE MR0175 or ISO
15156 (revision and materials of construction as specied by the
customer). Environmental restrictions may apply to wetted parts
and/or bolting.”
Page 89
Chemical Compatibility of Elastomers and Metals
659
Te c h n i c a l
Introduction
This section explains the uses and compatibilities of elastomers
commonly used in Fisher® regulators. The following tables
provide the compatibility of the most common elastomers and
metals to a variety of chemicals and/or compounds.
The information contained herein is extracted from data we believe
to be reliable. However, because of variable service conditions over
which we have no control, we do not in any way make any warranty,
either express or implied, as to the properties of any materials or as to
the performance of any such materials in any particular application, and
we hereby expressly disclaim any responsibility for the accuracy of any
of the information set forth herein.
Refer to the applicable process gas service code or standard
to determine if a specic material found in the Process Gases
Application Guide is allowed to be used in that service.
Elastomers: Chemical Names and Uses
NBR - Nitrile Rubber, also called Buna-N, is a copolymer of
butadiene and acrylonitrile. Nitrile is recommended for: general
purpose sealing, petroleum oils and uids, water, silicone greases
and oils, di-ester based lubricants (such as MIL-L-7808), and
ethylene glycol based uids (Hydrolubes). It is not recommended
for: halogenated hydrocarbons, nitro hydrocarbons (such as
nitrobenzene and aniline), phosphate ester hydraulic uids
(Skydrol, Cellulube, Pydraul), ketones (MEK, acetone), strong
acids, ozone, and automotive brake uid. Its temperature range is
-60° to 225°F (-51° to 107°C), although this would involve more
than one compound and would depend upon the stress state of the
component in service.
EPDM, EPM - Ethylenepropylene rubber is an elastomer prepared
from ethylene and propylene monomers. EPM is a copolymer of
ethylene and propylene, while EPDM contains a small amount
of a third monomer (a diene) to aid in the curing process. EP is
recommended for: phosphate ester based hydraulic uids, steam to
400°F (204°C), water, silicone oils and greases, dilute acids, dilute
alkalis, ketones, alcohols, and automotive brake uids. It is not
recommended for: petroleum oils, and di-ester based lubricants.
Its temperature range is -60° to 500°F (-51° to 260°C) (The high
limit would make use of a special high temperature formulation
developed for geothermal applications).
FKM - This is a uoroelastomer of the polymethylene type having
substituent uoro and peruoroalkyl or peruoroalkoxy groups
on the polymer chain. Viton® and Fluorel® are the most common
trade names. FKM is recommended for: petroleum oils, di-ester
based lubricants, silicate ester based lubricants (such as MLO
8200, MLO 8515, OS-45), silicone uids and greases, halogenated
hydrocarbons, selected phosphate ester uids, and some acids. It
is not recommended for: ketones, Skydrol 500, amines (UDMH),
anhydrous ammonia, low molecular weight esters and ethers, and
hot hydrouoric and chlorosulfonic acids. Its temperature range is
-20° to 450°F (-29° to 232°C) (This extended range would require
special grades and would limit use on each end of the range.).
CR - This is chloroprene, commonly know as neoprene, which
is a homopolymer of chloroprene (chlorobutadiene). CR is
recommended for: refrigerants (Freons, ammonia), high aniline
point petroleum oils, mild acids, and silicate ester uids. It is
not recommended for: phosphate ester uids and ketones. Its
temperature range is -60° to 200°F (-51° to 93°C), although this
would involve more than one compound.
NR - This is natural rubber which is a natural polyisoprene,
primarily from the tree, Hevea Brasiliensis. The synthetics
have all but completely replaced natural rubber for seal use.
NR is recommended for automotive brake uid, and it is not
recommended for petroleum products. Its temperature range is
-80° to 180°F (-62° to 82°C).
FXM - This is a copolymer of tetrauoroethylene and propylene;
hence, it is sometimes called PTFE/P rubber. Common trade
names are Aas® (Asahi Glass Co., Ltd) and Fluoraz® (Greene,
Tweed & Co.). It is generally used where resistance to both
hydrocarbons and hot water are required. Its temperature range is
20° to 400°F (-7° to 204°C).
ECO - This is commonly called Hydrin® rubber, although that is a
trade name for a series of rubber materials by B.F. Goodrich. CO
is the designation for the homopolymer of epichlorohydrin, ECO is
the designation for a copolymer of ethylene oxide and chloromethyl
oxirane (epichlorohydrin copolymer), and ETER is the designation
for the terpolymer of epichlorohydrin, ethylene oxide, and an
unsaturated monomer. All the epichlorohydrin rubbers exhibit
better heat resistance than nitrile rubbers, but corrosion with
aluminum may limit applications. Normal temperature range is
(-40° to 250°F (-40° to 121°C), while maximum temperature ranges
are -40° to 275°F (-40° to 135°C) (for homopolymer CO) and
-65° to 275°F (-54° to 135°C) (for copolymer ECO and
terpolymer ETER).
FFKM - This is a peruoroelastomer generally better known as
Kalrez® (DuPont) and Chemraz® (Greene, Tweed). Peruoro
rubbers of the polymethylene type have all substituent groups on
the polymer chain of uoro, peruoroalkyl, or peruoroalkoxy
groups. The resulting polymer has superior chemical resistance
and heat temperature resistance. This elastomer is extremely
expensive and should be used only when all else fails. Its
temperature range is 0° to 480°F (-18° to 249°C). Some materials,
such as Kalrez® 1050LF is usable to 550°F (288°C) and
Kalrez® 4079 can be used to 600°F (316°C).
FVMQ - This is uorosilicone rubber which is an elastomer that
should be used for static seals because it has poor mechanical
properties. It has good low and high temperature resistance and
is reasonably resistant to oils and fuels because of its uorination.
Because of the cost, it only nds specialty use. Its temperature
range is -80° to 400°F (-62° to 204°C).
VMQ - This is the most general term for silicone rubber. Silicone
rubber can be designated MQ, PMQ, and PVMQ, where the Q
designates any rubber with silicon and oxygen in the polymer
chain, and M, P, and V represent methyl, phenyl, and vinyl
substituent groups on the polymer chain. This elastomer is used
only for static seals due to its poor mechanical properties. Its
temperature range is -175° to 600°F (-115° to 316°C) (Extended
temperature ranges require special compounds for high or
low temperatures).
Page 90
Chemical Compatibility of Elastomers and Metals
Te c h n i c a l
660
General Properties of Elastomers
PROPERTY
NATURAL
RUBBER
BUNA-S
NITRILE
(NBR)
NEO-
PRENE
(CR)
BUTYL THIOKOL®SILICONE HYPALON
®
FLUORO-
ELASTOMER
(1,2)
(FKM)
POLY-
URETHANE
(2)
POLY-
ACRYLIC
(1)
ETHYLENE-
PROPYLENE
( 3)
(EPDM)
Tensile
Strength,
Psi (bar)
Pure Gum
3000
(207)
400
(28)
600
(41)
3500
(241)
3000
(207)
300
(21)
200 to 450
(14 to 31)
4000
(276)
- - - - - - - -
100
(7)
- - - -
Reinforced
4500
(310)
3000 (207)
4000
(276)
3500
(241)
3000
(207)
1500
(103)
1100
(76)
4400
(303)
2300
(159)
6500
(448)
1800
(124)
2500
(172)
Tear Resistance Excellent Poor-Fair Fair Good Good Fair Poor-Fair Excellent Good Excellent Fair Poor
Abrasion Resistance Excellent Good Good Excellent Fair Poor Poor Excellent Very Good Excellent Good Good
Aging: Sunlight
Oxidation
Poor
Good
Poor
Fair
Poor
Fair
Excellent
Good
Excellent
Good
Good
Good
Good
Very Good
Excellent
Very Good
Excellent
Excellent
Excellent
Excellent
Excellent
Excellent Good
Heat
(Maximum
Temperature)
200°F
(93°C)
200°F
(93°C)
250°F
(121°C)
200°F
(93°C)
200°F
(93°C)
140°F
(60°C)
450°F
(232°C)
300°F
(149°C)
400°F
(204°C)
200°F
(93°C)
350°F
(177°C)
350°F
(177°C)
Static (Shelf) Good Good Good Very Good Good Fair Good Good - - - - - - - - Good Good
Flex
Cracking Resistance
Excellent Good Good Excellent Excellent Fair Fair Excellent - - - - Excellent Good - - - -
Compression Set
Resistance
Good Good
Very
Good
Excellent Fair Poor Good Poor Poor Good Good Fair
Solvent Resistance:
Aliphatic Hydrocarbon
Aromatic Hydrocarbon
Oxygenated Solvent
Halogenated Solvent
Very Poor
Very Poor
Good
Very Poor
Very Poor
Very Poor
Good
Very Poor
Good
Fair
Poor
Very Poor
Fair
Poor
Fair
Very Poor
Poor
Very Poor
Good
Poor
Excellent
Good
Fair
Poor
Poor
Very Poor
Poor
Very Poor
Fair
Poor
Poor
Very Poor
Excellent
Very Good
Good
- - - -
Very Good
Fair
Poor
- - - -
Good
Poor
Poor
Poor
Poor
Fair
- - - Poor
Oil Resistance:
Low Aniline Mineral Oil
High Aniline Mineral
Oil
Synthetic Lubricants
Organic Phosphates
Very Poor
Very Poor
Very Poor
Very Poor
Very Poor
Very Poor
Very Poor
Very Poor
Excellent
Excellent
Fair
Very Poor
Fair
Good
Very Poor
Very Poor
Very Poor
Very Poor
Poor
Good
Excellent
Excellent
Poor
Poor
Poor
Good
Fair
Poor
Fair
Good
Poor
Poor
Excellent
Excellent
- - - Poor
- - - -
- - - -
- - - Poor
Excellent
Excellent
Fair
Poor
Poor
Poor
Poor
Very Good
Gasoline Resistance:
Aromatic
Non-Aromatic
Very Poor
Very Poor
Very Poor
Very Poor
Good
Excellent
Poor
Good
Very Poor
Very Poor
Excellent
Excellent
Poor
Good
Poor
Fair
Good
Very Good
Fair
Good
Fair
Poor
Fair
Poor
Acid Resistance:
Diluted (Under 10%)
Concentrated
(4)
Good
Fair
Good
Poor
Good
Poor
Fair
Fair
Good
Fair
Poor
Very Poor
Fair
Poor
Good
Good
Excellent
Very Good
Fair
Poor
Poor
Poor
Very Good
Good
Low Temperature
Flexibility (Maximum)
-65°F
(-54°C)
-50°F
(-46°C)
-40°F
(-40°C)
-40°F
(-40°C)
-40°F
(-40°C)
-40°F
(-40°C)
-100°F
(-73°C)
-20°F
(-29°C)
-30°F
(-34°C)
-40°F
(-40°C)
-10°F
(-23°C)
-50°F
(-45°C)
Permeability to Gases Fair Fair Fair Very Good Very Good Good Fair Very Good Good Good Good Good
Water Resistance Good Very Good
Very
Good
Fair Very Good Fair Fair Fair Excellent Fair Fair Very Good
Alkali Resistance:
Diluted (Under 10%)
Concentrated
Good
Fair
Good
Fair
Good
Fair
Good
Good
Very Good
Very Good
Poor
Poor
Fair
Poor
Good
Good
Excellent
Very Good
Fair
Poor
Poor
Poor
Excellent
Good
Resilience Very Good Fair Fair Very Good Very Good Poor Good Good Good Fair Very Poor Very Good
Elongation (Maximum) 700% 500% 500% 500% 700% 400% 300% 300% 425% 625% 200% 500%
1. Do not use with steam.
2. Do not use with ammonia.
3. Do not use with petroleum based uids. Use with ester based non-ammable hydraulic oils and low pressure steam applications to 300°F (149°C).
4. Except for nitric and sulfuric acid.
Page 91
Chemical Compatibility of Elastomers and Metals
661
Te c h n i c a l
Fluid Compatibility of Elastomers
FLUID
MATERIAL
Neoprene (CR) Nitrile (NBR) Fluoroelastomer (FKM)
Ethylenepropylene
(EPDM)
Peruoroelastomer
(FFKM)
Acetic Acid (30%)
Acetone
Air, Ambient
Air, Hot (200°F (93°C))
Alcohol (Ethyl)
Alcohol (Methyl)
Ammonia (Anhydrous) (Cold)
B
C
A
C
A
A
A
C
C
A
B
C
A
A
C
C
A
A
C
C
C
A
A
A
A
A
A
A
A
A
A
A
A
A
A
Ammonia (Gas, Hot)
Beer
Benzene
Brine (Calcium Chloride)
Butadiene Gas
Butane (Gas)
B
A
C
A
C
A
C
A
C
A
C
A
C
A
B
B
B
A
B
A
C
A
C
C
A
A
A
A
A
A
Butane (Liquid)
Carbon Tetrachloride
Chlorine (Dry)
Chlorine (Wet)
Coke Oven Gas
C
C
C
C
C
A
C
C
C
C
A
A
A
B
A
C
C
C
C
C
A
A
A
A
A
Ethyl Acetate
Ethylene Glycol
Freon 11
Freon 12
Freon 22
C
A
C
A
A
C
A
B
A
C
C
A
A
B
C
B
A
C
B
A
A
A
A
A
A
Freon 114
Gasoline (Automotive)
Hydrogen Gas
Hydrogen Sulde (Dry)
Hydrogen Sulde (Wet)
A
C
A
A
B
A
B
A
A
(1)
C
B
A
A
C
C
A
C
A
A
A
A
A
A
A
A
Jet Fuel (JP-4)
Methyl Ethyl Ketone (MEK)
MTBE
Natural Gas
B
C
C
A
A
C
C
A
A
C
C
A
C
A
C
C
A
A
A
A
Nitric Acid (50 to 100%)
Nitrogen
Oil (Fuel)
Propane
C
A
C
B
C
A
A
A
B
A
A
A
C
A
C
C
A
A
A
A
Sulfur Dioxide
Sulfuric Acid (up to 50%)
Sulfuric Acid (50 to 100%)
Water (Ambient)
Water (at 200°F (93°C))
A
B
C
A
C
C
C
C
A
B
A
A
A
A
B
A
B
B
A
A
A
A
A
A
A
1. Performance worsens with hot temperatures.
A - Recommended
B - Minor to moderate effect. Proceed with caution.
C - Unsatisfactory
N/A - Information not available
Page 92
Chemical Compatibility of Elastomers and Metals
Te c h n i c a l
662
- continued -
Compatibility of Metals
CORROSION INFORMATION
Fluid
Material
Carbon
Steel
Cast
Iron
S302
or S304
Stainless
Steel
S316
Stainless
Steel
Bronze Monel
®
Hastelloy
®
B
Hastelloy®
C
Durimet®
20
Titanium
Cobalt-
Base
Alloy 6
S416
Stainless
Steel
440C
Stainless
Steel
17-4PH
Stainless
Steel
Acetaldehyde
Acetic Acid, Air Free
Acetic Acid, Aerated
Acetic Acid Vapors
Acetone
A
C
C
C
A
A
C
C
C
A
A
B
A
A
A
A
B
A
A
A
A
B
A
B
A
A
B
A
B
A
IL
A
A
IL
A
A
A
A
A
A
A
A
A
B
A
IL
A
A
A
A
IL
A
A
A
A
A
C
C
C
A
A
C
C
C
A
A
B
B
B
A
Acetylene
Alcohols
Aluminum Sulfate
Ammonia
Ammonium Chloride
A
A
C
A
C
A
A
C
A
C
A
A
A
A
B
A
A
A
A
B
IL
A
B
C
B
A
A
B
A
B
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
IL
A
A
A
A
A
A
IL
A
B
A
A
C
A
C
A
A
C
A
C
A
A
IL
IL
IL
Ammonium Nitrate
Ammonium Phosphate
(Mono Basic)
Ammonium Sulfate
Ammonium Sulte
Aniline
A
C
C
C
C
C
C
C
C
C
A
A
B
A
A
A
A
A
A
A
C
B
B
C
C
C
B
A
C
B
A
A
A
IL
A
A
A
A
A
A
A
B
A
A
A
A
A
A
A
A
A
A
A
A
A
C
B
C
B
C
B
B
C
B
C
IL
IL
IL
IL
IL
Asphalt
Beer
Benzene (Benzol)
Benzoic Acid
Boric Acid
A
B
A
C
C
A
B
A
C
C
A
A
A
A
A
A
A
A
A
A
A
B
A
A
A
A
A
A
A
A
A
A
A
IL
A
A
A
A
A
A
A
A
A
A
A
IL
A
A
A
A
A
A
A
IL
A
A
B
A
A
B
A
B
A
A
B
A
A
A
A
IL
Butane
Calcium Chloride
(Alkaline)
Calcium Hypochlorite
Carbolic Acid
Carbon Dioxide, Dry
A
B
C
B
A
A
B
C
B
A
A
C
B
A
A
A
B
B
A
A
A
C
B
A
A
A
A
B
A
A
A
A
C
A
A
A
A
A
A
A
A
A
A
A
A
IL
A
A
A
A
A
IL
IL
A
A
A
C
C
IL
A
A
C
C
IL
A
A
IL
IL
IL
A
Carbon Dioxide, Wet
Carbon Disulde
Carbon Tetrachloride
Carbonic Acid
Chlorine Gas, Dry
C
A
B
C
A
C
A
B
C
A
A
A
B
B
B
A
A
B
B
B
B
C
A
B
B
A
B
A
A
A
A
A
B
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
IL
C
A
A
IL
IL
B
A
B
C
A
C
A
B
A
A
C
A
IL
IL
A
C
Chlorine Gas, Wet
Chlorine, Liquid
Chromic Acid
Citric Acid
Coke Oven Gas
C
C
C
IL
A
C
C
C
C
A
C
C
C
B
A
C
C
B
A
A
C
B
C
A
B
C
C
A
B
B
C
C
C
A
A
B
A
A
A
A
C
B
C
A
A
A
C
A
A
A
B
B
B
IL
A
C
C
C
B
A
C
C
C
B
A
C
C
C
B
A
Copper Sulfate
Cottonseed Oil
Creosote
Ethane
Ether
C
A
A
A
B
C
A
A
A
B
B
A
A
A
A
B
A
A
A
A
B
A
C
A
A
C
A
A
A
A
IL
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
IL
A
A
IL
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
Ethyl Chloride
Ethylene
Ethylene Glycol
Ferric Chloride
Formaldehyde
C
A
A
C
B
C
A
A
C
B
A
A
A
C
A
A
A
A
C
A
A
A
A
C
A
A
A
A
C
A
A
A
IL
C
A
A
A
IL
B
A
A
A
A
C
A
A
A
IL
A
A
A
A
A
B
A
B
A
A
C
A
B
A
A
C
A
IL
A
A
IL
A
Formic Acid
Freon, Wet
Freon, Dry
Furfural
Gasoline, Rene
IL
B
B
A
A
C
B
B
A
A
B
B
A
A
A
B
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
C
A
A
A
A
B
A
A
A
A
C
IL
IL
B
A
C
IL
IL
B
A
B
IL
IL
IL
A
A - Recommended
B - Minor to moderate effect. Proceed with caution.
C - Unsatisfactory
IL - Information lacking
Page 93
Chemical Compatibility of Elastomers and Metals
663
Te c h n i c a l
Compatibility of Metals (continued)
CORROSION INFORMATION
Fluid
Material
Carbon
Steel
Cast
Iron
S302
or S304
Stainless
Steel
S316
Stainless
Steel
Bronze Monel
®
Hastelloy
®
B
Hastelloy® CDurimet®
20
Titanium
Cobalt-
Base
Alloy 6
S416
Stainless
Steel
440C
Stainless
Steel
17-4PH
Stainless
Steel
Glucose
Hydrochloric Acid, Aerated
Hydrochloric Acid, Air free
Hydrouoric Acid, Aerated
Hydrouoric Acid, Air free
A
C
C
B
A
A
C
C
C
C
A
C
C
C
C
A
C
C
B
B
A
C
C
C
C
A
C
C
C
A
A
A
A
A
A
A
B
B
A
A
A
C
C
B
B
A
C
C
C
C
A
B
B
B
IL
A
C
C
C
C
A
C
C
C
C
A
C
C
C
IL
Hydrogen
Hydrogen Peroxide
Hydrogen Sude, Liquid
Magnesium Hydroxide
Mercury
A
IL
C
A
A
A
A
C
A
A
A
A
A
A
A
A
A
A
A
A
A
C
C
B
C
A
A
C
A
B
A
B
A
A
A
A
B
A
A
A
A
A
B
A
A
A
A
A
A
A
A
IL
A
A
A
A
B
C
A
A
A
B
C
A
A
A
IL
IL
IL
B
Methanol
Methyl Ethyl Ketone
Milk
Natural Gas
Nitric Acid
A
A
C
A
C
A
A
C
A
C
A
A
A
A
A
A
A
A
A
B
A
A
A
A
C
A
A
A
A
C
A
A
A
A
C
A
A
A
A
B
A
A
A
A
A
A
IL
A
A
A
A
A
A
A
C
A
A
C
A
C
B
A
C
A
C
A
A
C
A
B
Oleic Acid
Oxalic Acid
Oxygen
Petroleum Oils, Rened
Phosphoric Acid, Aerated
C
C
A
A
C
C
C
A
A
C
A
B
A
A
A
A
B
A
A
A
B
B
A
A
C
A
B
A
A
C
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
B
A
A
B
A
B
A
A
A
A
B
A
A
C
A
B
A
A
C
IL
IL
A
A
IL
Phosphoric Acid, Air Free
Phosphoric Acid Vapors
Picric Acid
Potassium Chloride
Potassium Hydroxide
C
C
C
B
B
C
C
C
B
B
A
B
A
A
A
A
B
A
A
A
C
C
C
B
B
B
C
C
B
A
A
A
A
A
A
A
IL
A
A
A
A
A
A
A
A
B
B
IL
A
A
A
C
IL
IL
IL
C
C
B
C
B
C
C
B
C
B
IL
IL
IL
IL
IL
Propane
Rosin
Silver Nitrate
Sodium Acetate
Sodium Carbonate
A
B
C
A
A
A
B
C
A
A
A
A
A
B
A
A
A
A
A
A
A
A
C
A
A
A
A
C
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
IL
A
A
A
A
A
B
A
A
A
A
B
A
B
A
A
B
A
B
A
A
IL
A
A
Sodium Chloride
Sodium Chromate
Sodium Hydroxide
Sodium Hypochloride
Sodium Thiosulfate
C
A
A
C
C
C
A
A
C
C
B
A
A
C
A
B
A
A
C
A
A
A
C
B-C
C
A
A
A
B-C
C
A
A
A
C
A
A
A
A
A
A
A
A
A
B
A
A
A
A
A
A
A
A
A
IL
IL
B
A
B
C
B
B
A
B
C
B
B
A
A
IL
IL
Stannous Chloride
Stearic Acid
Sulfate Liquor (Black)
Sulfur
Sulfur Dioxide, Dry
B
A
A
A
A
B
C
A
A
A
C
A
A
A
A
A
A
A
A
A
C
B
C
C
A
B
B
A
A
A
A
A
A
A
B
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
IL
B
A
A
A
C
B
IL
A
B
C
B
IL
A
B
IL
IL
IL
A
IL
Sulfur Trioxide, Dry
Sufuric Acid (Aerated)
Sufuric Acid (Air Free)
Sulfurous Acid
Tar
A
C
C
C
A
A
C
C
C
A
A
C
C
B
A
A
C
C
B
A
A
C
B
B
A
A
C
B
C
A
B
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
B
B
A
A
A
B
B
B
A
B
C
C
C
A
B
C
C
C
A
IL
C
C
IL
A
Trichloroethylene
Turpentine
Vinegar
Water, Boiler Feed
Water, Distilled
B
B
C
B
A
B
B
C
C
A
B
A
A
A
A
A
A
A
A
A
A
A
B
C
A
A
B
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
IL
A
A
A
A
A
A
A
B
A
C
B
B
B
A
C
A
B
IL
A
A
A
IL
Water, Sea
Whiskey and Wines
Zinc Chloride
Zinc Sulfate
B
C
C
C
B
C
C
C
B
A
C
A
B
A
C
A
A
A
C
B
A
B
C
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
A
B
A
C
C
C
B
C
C
C
B
A
IL
IL
IL
A - Recommended
B - Minor to moderate effect. Proceed with caution.
C - Unsatisfactory
IL - Information lacking
Page 94
Regulator Tips
664
Te c h n i c a l
1. All regulators should be installed and used in accordance with
federal, state, and local codes and regulations.
2. Adequate overpressure protection should be installed to protect
the regulator from overpressure. Adequate overpressure
protection should also be installed to protect all downstream
equipment in the event of regulator failure.
3. Downstream pressures signicantly higher than the regulator's
pressure setting may damage soft seats and
other internal parts.
4. If two or more available springs have published pressure
ranges that include the desired pressure setting, use the spring
with the lower range for better accuracy.
5. The recommended selection for orice diameters is the
smallest orice that will handle the ow.
6. Most regulators shown in this application guide are generally
suitable for temperatures to 180°F (82°C). With high
temperature uoroelastomers (if available), the regulators
can be used for temperatures to 300°F (149°C). Check
the temperature capabilities to determine materials and
temperature ranges available. Use stainless steel diaphragms
and seats for higher temperatures, such as steam service.
7. The full advertised range of a spring can be utilized without
sacricing performance or spring life.
8. Regulator body size should not be larger than the pipe size. In
many cases, the regulator body is one size smaller than the
pipe size.
9. Do not oversize regulators. Pick the smallest orice size or
regulator that will work. Keep in mind when sizing a station
that most restricted trims that do not reduce the main port size
do not help with improved low ow control.
10. Speed of regulator response, in order:
• Direct-operated
• Two-path pilot-operated
• Unloading pilot-operated
• Control valve
Note: Although direct-operated regulators give the fastest
response, all types provide quick response.
11. When a regulator appears unable to pass the published
ow rate, be sure to check the inlet pressure measured at
the regulator body inlet connection. Piping up to and away
from regulators can cause signicant owing pressure losses.
12. When adjusting setpoint, the regulator should be owing at
least ve percent of the normal operating ow.
13. Direct-operated regulators generally have faster response to
quick ow changes than pilot-operated regulators.
14. Droop is the reduction of outlet pressure experienced by
pressure-reducing regulators as the ow rate increases. It is
stated as a percent, in inches of water column (mbar) or in
pounds per square inch (bar) and indicates the difference
between the outlet pressure setting made at low ow rates
and the actual outlet pressure at the published maximum
ow rate. Droop is also called offset or proportional band.
15. Downstream pressure always changes to some extent when
inlet pressure changes.
16. Most soft-seated regulators will maintain the pressure within
reasonable limits down to zero ow. Therefore, a regulator
sized for a high ow rate will usually have a turndown ratio
sufcient to handle pilot-light loads during off cycles.
17. Do not undersize the monitor set. It is important to realize
that the monitor regulator, even though it is wide-open,
will require pressure drop for ow. Using two identical
regulators in a monitor set will yield approximately
70 percent of the capacity of a single regulator.
18. Diaphragms leak a small amount due to migration of gas
through the diaphragm material. To allow escape of this gas,
be sure casing vents (where provided) remain open.
19. Use control lines of equal or greater size than the control tap
on the regulator. If a long control line is required, make it
bigger. A rule of thumb is to use the next nominal pipe size
for every 20 feet (6,1 m) of control line. Small control
lines cause a delayed response of the regulator, leading
to increased chance of instability. 3/8-inch (9,5 mm) OD
tubing is the minimum recommended control line size.
20. For every 15 psid (1,0 bar d) pressure differential
across the regulator, expect approximately a one degree
drop in gas temperature due to the natural refrigeration
effect. Freezing is often a problem when the ambient
temperature is between 30° and 45°F (-1° and 7°C).
21. A disk with a cookie cut appearance probably means you had
an overpressure situation. Thus, investigate further.
22. When using relief valves, be sure to remember that the
reseat point is lower than the start-to-bubble point. To
avoid seepage, keep the relief valve setpoint far enough
above the regulator setpoint.
Page 95
Regulator Tips
665
Te c h n i c a l
23. Vents should be pointed down to help avoid the accumulation
of water condensation or other materials in the spring case.
24. Make control line connections in a straight run of pipe about
10 pipe diameters downstream of any area of turbulence,
such as elbows, pipe swages, or block valves.
25. When installing a working monitor station, get as much
volume between the two regulators as possible. This
will give the upstream regulator more room to control
intermediate pressure.
26. Cutting the supply pressure to a pilot-operated regulator
reduces the regulator gain or sensitivity and, thus, may
improve regulator stability. (This can only be used with two
path control.)
27. Regulators with high ows and large pressure drops generate
noise. Noise can wear parts which can cause failure and/or
inaccurate control. Keep regulator noise below 110 dBA.
28. Do not place control lines immediately downstream of rotary
or turbine meters.
29. Keep vents open. Do not use small diameter, long vent lines.
Use the rule of thumb of the next nominal pipe size every
10 feet (3,1 m) of vent line and 3 feet (0,9 m) of vent line
for every elbow in the line.
30. Fixed factor measurement (or PFM) requires the regulator
to maintain outlet pressure within ±1% of absolute pressure.
For example: Setpoint of 2 psig + 14.7 psia = 16.7 psia x
0.01 = ±0.167 psi. (Setpoint of 0,14 bar + 1,01 bar = 1,15 bar x
0,01 = ±0,0115 bar.)
31. Regulating Cg (coefcient of ow) can only be used for
calculating ow capacities on pilot-operated regulators.
Use capacity tables or ow charts for determining a direct-
operated regulator’s capacity.
32. Do not make the setpoints of the regulator/monitor too close
together. The monitor can try to take over if the setpoints are
too close, causing instability and reduction of capacity. Set
them at least one proportional band apart.
33. Consider a butt-weld end regulator where available to lower
costs and minimize ange leakages.
34. Do not use needle valves in control lines; use full-open
valves. Needle valves can cause instability.
35. Burying regulators is not recommended. However, if you
must, the vent should be protected from ground moisture
and plugging.
Page 96
666
Te c h n i c a l
Conversions, Equivalents, and Physical Data
Pressure Equivalents
KG PER
SQUARE
CENTIMETER
POUNDS PER
SQUARE INCH
ATMOSPHERE BAR
INCHES OF
MERCURY
KILOPASCALS
INCHES OF
WATER COLUMN
FEET OF
WATER COLUMN
Kg per square cm 1 14.22 0.9678 0,98067 28.96 98,067 394.05 32.84
Pounds per square inch 0,07031 1 0.06804 0,06895 2.036 6,895 27.7 2.309
Atmosphere 1,0332 14.696 1 1,01325 29.92 101,325 407.14 33.93
Bar 1,01972 14.5038 0.98692 1 29.53 100 402.156 33.513
Inches of Mercury 0,03453 0.4912 0.03342 0,033864 1 3,3864 13.61 1.134
Kilopascals 0,0101972 0.145038 0.0098696 0,01 0.2953 1 4.02156 0.33513
Inches of Water 0,002538 0.0361 0.002456 0,00249 0.07349 0,249 1 0.0833
Feet of Water 0,3045 0.4332 0.02947 0,029839 0.8819 2,9839 12 1
1 ounce per square inch = 0.0625 pounds per square inch
BY
TO
OBTAIN
MULTIPLY
NUMBER
OF
Volume Equivalents
CUBIC
DECIMETERS
(LITERS)
CUBIC INCHES CUBIC FEET U.S. QUART U.S. GALLON IMPERIAL GALLON
U.S. BARREL
(PETROLEUM)
Cubic Decimeters (Liters) 1 61.0234 0.03531 1.05668 0.264178 0,220083 0.00629
Cubic Inches 0,01639 1 5.787 x 10
-4
1.01732 0.004329 0,003606 0.000103
Cubic Feet 28,317 1728 1 29.9221 7.48055 6,22888 0.1781
U.S. Quart 0,94636 57.75 0.03342 1 0.25 0,2082 0.00595
U.S. Gallon 3,78543 231 0.13368 4 1 0,833 0.02381
Imperial Gallon 4,54374 277.274 0.16054 4.80128 1.20032 1 0.02877
U.S. Barrel (Petroleum) 158,98 9702 5.6146 168 42 34,973 1
1 cubic meter = 1,000,000 cubic centimeters
1 liter = 1000 milliliters = 1000 cubic centimeters
BY
TO
OBTAIN
MULTIPLY
NUMBER
OF
Pressure Conversion - Pounds per Square Inch to Bar
(1)
POUNDS PER
SQUARE INCH
0 1 2 3 4 5 6 7 8 9
Bar
0
10
20
30
0,000
0,689
1,379
2,068
0,069
0,758
1,448
2,137
0,138
0,827
1,517
2,206
0,207
0,896
1,586
2,275
0,276
0,965
1,655
2,344
0,345
1,034
1,724*
2,413
0,414
1,103
1,793
2,482
0,482
1,172
1,862
2,551
0,552
1,241
1,931
2,620
0,621
1,310
1,999
2,689
40
50
60
70
2,758
3,447
4,137
4,826
2,827
3,516
4,275
4,964
2,896
3,585
4,275
4,964
2,965
3,654
4,344
5,033
3,034
3,723
4,413
5,102
3,103
3,792
4,482
5,171
3,172
3,861
4,551
5,240
3,241
3,930
4,619
5,309
3,309
3,999
4,688
5,378
3,378
4,068
4,758
5,447
80
90
100
5,516
6,205
6,895
5,585
6,274
6,964
5,654
6,343
7,033
5,723
6,412
7,102
5,792
6,481
7,171
5,861
6,550
7,239
5,929
6,619
7,308
5,998
6,688
7,377
6,067
6,757
7,446
6,136
6,826
7,515
1. To convert to kilopascals, move decimal point two positions to the right; to convert to megapascals, move decimal point one position to the left.
*Note: Round off decimal points to provide no more than the desired degree of accuracy.
To use this table, see the shaded example.
25 psig (20 from the left column plus ve from the top row) = 1,724 bar
Page 97
667
Te c h n i c a l
Conversions, Equivalents, and Physical Data
Volume Rate Equivalents
LITERS
PER MINUTE
CUBIC METERS
PER HOUR
CUBIC FEET
PER HOUR
LITERS
PER HOUR
U.S. GALLONS
PER MINUTE
U.S. BARRELS
PER DAY
Liters per Minute 1 0,06 2.1189 60 0.264178 9.057
Cubic Meters per Hour 16,667 1 35.314 1000 4.403 151
Cubic Feet per Hour 0,4719 0,028317 1 28.317 0.1247 4.2746
Liters per Hour 0,016667 0,001 0.035314 1 0.004403 0.151
U.S. Gallons per Minute 3,785 0,2273 8.0208 227.3 1 34.28
U.S. Barrels per Day 0,1104 0,006624 0.23394 6.624 0.02917 1
BY
TO
OBTAIN
MULTIPLY
NUMBER
OF
Temperature Conversion Formulas
TO CONVERT FROM TO SUBSTITUTE IN FORMULA
Degrees Celsius Degrees Fahrenheit (°C x 9/5) + 32
Degrees Celsius Kelvin (°C + 273.16)
Degrees Fahrenheit Degrees Celsius (°F - 32) x 5/9
Degrees Fahrenheit Degrees Rankine (°F + 459.69)
Area Equivalents
SQUARE
METERS
SQUARE
INCHES
SQUARE
FEET
SQUARE
MILES
SQUARE
KILOMETERS
Square Meters 1 1549.99 10.7639 3.861 x 10-71 x 10
-6
Square Inches 0,0006452 1 6.944 x 10-32.491 x 10
-10
6,452 x 10
-10
Square Feet 0,0929 144 1 3.587 x 10-89,29 x 10
-8
Square Miles 2 589 999 - - - - 27,878,400 1 2,59
Square Kilometers 1 000 000 - - - - 10,763,867 0.3861 1
1 square meter = 10 000 square centimeters
1 square millimeter = 0,01 square centimeter = 0.00155 square inches
BY
TO
OBTAIN
MULTIPLY
NUMBER
OF
Kinematic-Viscosity Conversion Formulas
VISCOSITY SCALE RANGE OF t, SEC
KINEMATIC VISCOSITY,
STROKES
Saybolt Universal 32 < t < 100 t > 100
0.00226t - 1.95/t
0.00220t - 1.35/t
Saybolt Furol 25 < t < 40 t > 40
0.0224t - 1.84/t
0.0216t - 0.60/t
Redwood No. 1 34 < t < 100 t > 100
0.00226t - 1.79/t
0.00247t - 0.50/t
Redwood Admiralty - - - - 0.027t - 20/t
Engler - - - - 0.00147t - 3.74/t
Mass Conversion - Pounds to Kilograms
POUNDS
0 1 2 3 4 5 6 7 8 9
Kilograms
0
10
20
30
0,00
4,54
9,07
13,61
0,45
4,99
9,53
14,06
0,91
5,44
9,98
14,52
1,36
5,90
10,43
14,97
1,81
6,35
10,89
15,42
2,27
6,80
11,34*
15,88
2,72
7,26
11,79
16,33
3,18
7,71
12,25
16,78
3,63
8,16
12,70
17,24
4,08
8,62
13,15
17,69
40
50
60
70
18,14
22,68
27,22
31,75
18,60
23,13
27,67
32,21
19,05
23,59
28,12
32,66
19,50
24,04
28,58
33,11
19,96
24,49
29,03
33,57
20,41
24,95
29,48
34,02
20,87
25,40
29,94
34,47
21,32
25,86
30,39
34,93
21,77
26,31
30,84
35,38
22,23
26,76
31,30
35,83
809036,29
40,82
36,74
41,28
37,20
41,73
37,65
42,18
38,10
42,64
38,56
43,09
39,01
43,55
39,46
44,00
39,92
44,45
40,37
44,91
1 pound = 0,4536 kilograms
*NOTE: To use this table, see the shaded example.
25 pounds (20 from the left column plus ve from the top row) = 11,34 kilograms
Page 98
668
Te c h n i c a l
Conversions, Equivalents, and Physical Data
Conversion Units
MULTIPLY BY TO OBTAIN
Volume
Cubic centimeter 0.06103 Cubic inches
Cubic feet 7.4805 Gallons (US)
Cubic feet 28.316 Liters
Cubic feet 1728 Cubic inches
Gallons (US) 0.1337 Cubic feet
Gallons (US) 3.785 Liters
Gallons (US) 231 Cubic inches
Liters 1.057 Quarts (US)
Liters 2.113 Pints (US)
Miscellaneous
BTU 0.252 Calories
Decitherm 10,000 BTU
Kilogram 2.205 Pounds
Kilowatt Hour 3412 BTU
Ounces 28.35 Grams
Pounds 0.4536 Kilograms
Pounds 453.5924 Grams
Pounds 21,591 LPG BTU
Therm 100,000 BTU
API Bbls 42 Gallons (US)
Gallons of Propane 26.9 KWH
HP 746 KWH
HP (Steam) 42,418 BTU
Pressure
Grams per square centimeter 0.0142 Pounds per square inch
Inches of mercury 0.4912 Pounds per square inch
Inches of mercury 1.133 Feet of water
Inches of water 0.0361 Pounds per square inch
Inches of water 0.0735 Inches of mercury
Inches of water 0.5781 Ounces per square inch
Inches of water 5.204 Pounds per foot
kPa 100 Bar
Kilograms per square centimeter 14.22 Pounds per square inch
Kilograms per square meter 0.2048 Pounds per square foot
Pounds per square inch 0.06804 Atmospheres
Pounds per square inch 0.07031 Kilograms per square centimeter
Pounds per square inch 0.145 KPa
Pounds per square inch 2.036 Inches of mercury
Pounds per square inch 2.307 Feet of water
Pounds per square inch 14.5 Bar
Pounds per square inch 27.67 Inches of water
Length
Centimeters 0.3937 Inches
Feet 0.3048 Meters
Feet 30.48 Centimeters
Feet 304.8 Millimeters
Inches 2.540 Centimeters
Inches 25.40 Millimeters
Kilometer 0.6214 Miles
Meters 1.094 Yards
Meters 3.281 Feet
Meters 39.37 Inches
Miles (nautical) 1853 Meters
Miles (statute) 1609 Meters
Yards 0.9144 Meters
Yards 91.44 Centimeters
Other Useful Conversions
TO CONVERT FROM TO MULTIPLY BY
Cubic feet of methane BTU 1000 (approximate)
Cubic feet of water Pounds of water 62.4
Degrees Radians 0,01745
Gallons Pounds of water 8.336
Grams Ounces 0.0352
Horsepower (mechanical) Foot pounds per minute 33,000
Horsepower (electrical) Watts 746
Kg Pounds 2.205
Kg per cubic meter Pounds per cubic feet 0.06243
Kilowatts Horsepower 1.341
Pounds Kg 0,4536
Pounds of Air
(14.7 psia and 60°F)
Cubic feet of air 13.1
Pounds per cubic feet Kg per cubic meter 16,0184
Pounds per hour (gas) SCFH 13.1 ÷ Specic Gravity
Pounds per hour (water) Gallons per minute 0.002
Pounds per second (gas) SCFH 46,160 ÷ Specic Gravity
Radians Degrees 57.3
SCFH Air SCFH Propane 0.81
SCFH Air SCFH Butane 0.71
SCFH Air SCFH 0.6 Natural Gas 1.29
SCFH Cubic meters per hour 0.028317
Converting Volumes of Gas
CFH TO CFH OR CFM TO CFM
Multiply Flow of By To Obtain Flow of
Air
0.707 Butane
1.290 Natural Gas
0.808 Propane
Butane
1.414 Air
1.826 Natural Gas
1.140 Propane
Natural Gas
0.775 Air
0.547 Butane
0.625 Propane
Propane
1.237 Air
0.874 Butane
1.598 Natural Gas
Page 99
669
Te c h n i c a l
Conversions, Equivalents, and Physical Data
Length Equivalents
METERS INCHES FEET MILLIMETERS MILES KILOMETERS
Meters 1 39.37 3.2808 1000 0.0006214 0,001
Inches 0,0254 1 0.0833 25,4 0.00001578 0,0000254
Feet 0,3048 12 1 304,8 0.0001894 0,0003048
Millimeters 0,001 0.03937 0.0032808 1 0.0000006214 0,000001
Miles 1609,35 63,360 5,280 1 609 350 1 1,60935
Kilometers 1000 39,370 3280.83 1 000 000 0.62137 1
1 meter = 100 cm = 1000 mm = 0,001 km = 1,000,000 micrometers
BY
TO
OBTAIN
MULTIPLY
NUMBER
OF
Metric Prexes and Symbols
MULTIPLICATION FACTOR PREFIX SYMBOL
1 000 000 000 000 000 000 = 10
18
1 000 000 000 000 000 = 10
15
1 000 000 000 000 = 10
12
1 000 000 000 = 10
9
1 000 000 = 10
6
1 000 = 10
3
100 = 10
2
10 = 10
1
exa
peta
tera
giga
mega
kilo
hecto
deka
E
P
T
G
M
k
h
da
0.1 = 10
-1
0.01 = 10
-2
0.001 = 10
-3
0.000 01 = 10
-6
0.000 000 001 = 10
-9
0.000 000 000 001 = 10
-12
0.000 000 000 000 001 = 10
-15
0.000 000 000 000 000 001 = 10
-18
deci
centi
milli
micro
nano
pico
femto
atto
d
c
m
m
n
p
f
a
Greek Alphabet
CAPS
LOWER
CASE
GREEK
NAME
CAPS
LOWER
CASE
GREEK
NAME
CAPS
LOWER
CASE
GREEK
NAME
Α α Alpha Ι ι Iota Ρ ρ Rho
Β β Beta Κ κ Kappa Σ σ Sigma
Γ γ Gamma Λ λ Lambda Τ τ Tau
Δ δ Delta Μ μ Mu Υ υ Upsilon
Ε ε Epsilon Ν ν Nu Φ φ Phi
Ζ ζ Zeta Ξ ξ Xi Χ χ Chi
Η η Eta Ο ο Omicron Ψ ψ Psi
Θ θ Theta Π π Pi Ω ω Omega
Fractional Inches to Millimeters
INCH
0 1/16 1/8 3/16 1/4 5/16 3/8 7/16 1/2 9/16 5/8 11/16 3/4 13/16 7/8 15/16
mm
0
1
2
3
0,0
25,4
50,8
76,2
1,6
27,0
52,4
77,8
3,2
28,6
54,0
79,4
4,8
30,2
55,6
81,0
6,4
31,8
57,2
82,6
7,9
33,3
58,7
84,1
9,5
34,9
60,3
85,7
11,1
36,5
61,9
87,3
12,7
38,1
63,5
88,9
14,3
39,7
65,1
90,5
15,9
41,3
66,7
92,1
17,5
42,9
68,3
93,7
19,1
44,5
69,9
95,3
20,6
46,0
71,4
96,8
22,2
47,6
73,0
98,4
23,8
49,2
74,6
100,0
4
5
6
7
101,6
127,0
152,4
177,8
103,2
128,6
154,0
179,4
104,8
130,2
155,6
181,0
106,4
131,8
157,2
182,6
108,0
133,4
158,8
184,2
109,5
134,9
160,3
185,7
111,1
136,5
161,9
187,3
112,7
138,1
163,5
188,9
114,3
139,7
165,1
190,5
115,9
141,3
166,7
192,1
117,5
142,9
168,3
193,7
119,1
144,5
169,9
195,3
120,7
146,1
171,5
196,9
122,2
147,6
173,0
198,4
123,8
149,2
174,6
200,0
125,4
150,8
176,2
201,6
8
9
10
203,2
228,6
254,0
204,8
230,2
255,6
206,4
231,8
257,2
208,0
233,4
258,8
209,6
235,0
260,4
211,1
236,5
261,9
212,7
238,1
263,5
214,3
239,7
265,1
215,9
241,3
266,7
217,5
242,9
268,3
219,1
244,5
269,9
220,7
246,1
271,5
222,3
247,7
273,1
223,8
249,2
274,6
225,4
250,8
276,2
227,0
252,4
277,8
1-inch = 25,4 millimeters
NOTE: To use this table, see the shaded example.
2-1/2-inches (2 from the left column plus 1/2 from the top row) = 63,5 millimeters
Whole Inch-Millimeter Equivalents
INCH
0 1 2 3 4 5 6 7 8 9
mm
0 0,00 25,4 50,8 76,2 101,6 127,0 152,4 177,8 203,2 228,6
10 254,0 279,4 304,8 330,2 355,6 381,0 406,4 431,8 457,2 482,6
20 508,0 533,4 558,8 584,2 609,6 635,0 660,4 685,8 711,2 736,6
30 762,0 787,4 812,8 838,2 863,6 889,0 914,4 939,8 965,2 990,6
40 1016,0 1041,4 1066,8 1092,2 1117,6 1143,0 1168,4 1193,8 1219,2 1244,6
50 1270,0 1295,4 1320,8 1346,2 1371,6 1397,0 1422,4 1447,8 1473,2 1498,6
60 1524,0 1549,4 1574,8 1600,2 1625,6 1651,0 1676,4 1701,8 1727,2 1752,6
70 1778,0 1803,4 1828,8 1854,2 1879,6 1905,0 1930,4 1955,8 1981,2 2006,6
80 2032,0 2057,4 2082,8 2108,2 2133,6 2159,0 2184,4 2209,8 2235,2 2260,6
90 2286,0 2311,4 2336,8 2362,2 2387,6 2413,0 2438,4 2463,8 2489,2 2514,6
100 2540,0 2565,4 2590,8 2616,2 2641,6 2667,0 2692,4 2717,8 2743,2 2768,6
Note: All values in this table are exact, based on the relation 1-inch = 25,4 mm.
To use this table, see the shaded example.
25-inches (20 from the left column plus ve from the top row) = 635 millimeters
Page 100
670
Te c h n i c a l
Conversions, Equivalents, and Physical Data
Length Equivalents - Fractional and Decimal Inches to Millimeters
INCHES
mm
INCHES
mm
INCHES
mm
INCHES
mm
Fractions Decimals Fractions Decimals Fractions Decimals Fractions Decimals
0.00394 0.1 0.23 5.842 1/2 0.50 12.7 0.77 19.558
0.00787 0.2 15/64 0.234375 5.9531 0.51 12.954 0.78 19.812
0.01 0.254 0.23622 6.0 0.51181 13.0 25/32 0.78125 19.8438
0.01181 0.3 0.24 6.096 33/64 0.515625 13.0969 0.78740 20.0
1/64 0.015625 0.3969 1/4 0.25 6.35 0.52 13.208 0.79 20.066
0.01575 0.4 0.26 6.604 0.53 13.462 51/64 0.796875 20.2406
0.01969 0.5 17/64 0.265625 6.7469 17/32 0.53125 13.4938 0.80 20.320
0.02 0.508 0.27 6.858 0.54 13.716 0.81 20.574
0.02362 0.6 0.27559 7.0 35/64 0.546875 13.8906 13/64 0.8125 20.6375
0.02756 0.7 0.28 7.112 0.55 13.970 0.82 20.828
0.03 0.762 9/32 0.28125 7.1438 0.55118 14.0 0.82677 21.0
1/32 0.03125 0.7938 0.29 7.366 0.56 14.224 53/64 0.828125 21.0344
0.0315 0.8 19/64 0.296875 7.5406 9/16 0.5625 14.2875 0.83 21.082
0.13543 0.9 0.30 7.62 0.57 14.478 0.84 21.336
0.03937 1.0 0.31 7.874 37/64 0.578125 14.6844 27/32 0.84375 21.4312
0.04 1.016 5/16 0.3125 7.9375 0.58 14.732 0.85 21.590
3/64 0.046875 1.1906 0.31496 8.0 0.59 14.986 55/64 0.859375 21.8281
0.05 1.27 0.32 8.128 0.5905 15.0 0.86 21.844
0.06 1.524 21/64 0.328125 8.3344 19/32 0.59375 15.0812 0.86614 22.0
1/16 0.0625 1.5875 0.33 8.382 0.60 15.24 0.87 22.098
0.07 1.778 0.34 8.636 39/64 0.609375 15.4781 7/8 0.875 22.225
5/64 0.078125 1.9844 11/32 0.34375 8.7312 0.61 15.494 0.88 22.352
0.07874 2.0 0.35 8.89 0.62 15.748 0.89 22.606
0.08 2.032 0.35433 9.0 5/8 0.625 15.875 57/64 0.890625 22.6219
0.09 2.286 23/64 0.359375 9.1281 0.62992 16.0 0.90 22.860
3/32 0.09375 2.3812 0.36 9.144 0.63 16.002 0.90551 23.0
0.1 2.54 0.37 9.398 0.64 16.256 29/32 0.90625 23.0188
7/64 0.109375 2.7781 3/8 0.375 9.525 41/64 0.640625 16.2719 0.91 23.114
0.11 2.794 0.38 9.652 0.65 16.510 0.92 23.368
0.11811 3.0 0.39 9.906 21/32 0.65625 16.6688 59/64 0.921875 23.1456
0.12 3.048 25/64 0.390625 9.9219 0.66 16.764 0.93 23.622
1/8 0.125 3.175 0.39370 10.0 0.66929 17.0 15/16 0.9375 23.8125
0.13 3.302 0.40 10.16 0.67 17.018 0.94 23.876
0.14 3.556 13/32 0.40625 10.3188 43/64 0.671875 17.0656 0.94488 24.0
9/64 0.140625 3.5719 0.41 10.414 0.68 17.272 0.95 24.130
0.15 3.810 0.42 10.668 11/16 0.6875 17.4625 61/64 0.953125 24.2094
5/32 0.15625 3.9688 27/64 0.421875 10.7156 0.69 17.526 0.96 24.384
0.15748 4.0 0.43 10.922 0.70 17.78 31/32 0.96875 24.6062
0.16 4.064 0.43307 11.0 45/64 0.703125 17.8594 0.97 24.638
0.17 4.318 7/16 0.4375 11.1125 0.70866 18.0 0.98 24.892
11/64 0.171875 4.3656 0.44 11.176 0.71 18.034 0.98425 25.0
0.18 4.572 0.45 11.430 23/32 0.71875 18.2562 63/64 0.984375 25.0031
3/16 0.1875 4.7625 29/64 0.453125 11.5094 0.72 18.288 0.99 25.146
0.19 4.826 0.46 11.684 0.73 18.542 1 1.00000 25.4000
0.19685 5.0 15/32 0.46875 11.9062 47/64 0.734375 18.6531
0.2 5.08 0.47 11.938 0.74 18.796
13/64 0.203125 5.1594 0.47244 12.0 0.74803 19.0
0.21 5.334 0.48 12.192 3/4 0.75 19.050
7/32 0.21875 5.5562 31/64 0.484375 12.3031 0.76 19.304
0.22 5.588 0.49 12.446 49/64 0.765625 19.4469
Note: Round off decimal points to provide no more than the desired degree of accuracy.
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