EPRI NP-5067 Good Bolting Practices

GOOD· BOLTING PRACTICES

A Reference Manual for Nuclear power Plant Maintenance Personnel VOLUME 1: LARGE BOLT MANUAL

Prepared by JOHN H. BICKFORD Raymond Engineering Inc. 217 Smith Street Middletown, Connecticut 06457 MICHAEL E. LOORAM Looram Engineering 515 Main Street Yalesville, Connecticut 06492

EPRI Electric Power Research Institute 3412 Hillview Avenue Palo Alto, California 94304

TED MARSTON EPRI Project Manager

PRJ NP-5067

EPRI Electric Power Research Institute

GOOD BOLTING PRACTICES

A Reference Manual

--

for Nuclear Power Plant Maintena~e Personnel VOLUME 1: LARGE BOLT MANUAL

Prepared by JOHN BICKFORD Raymond Engineering Inc.

MICHAEL LOORAM Looram Engineering

ORDERING INFORMATION

Requests for copies of this report should be directed to Research Reports Center (RRC), Box 50490, Palo Alto, CA 94303, (415) 965-4081. There is no charge for reports requested by EPRI member utilities and affiliates, U.S,. utility associations, U.S. government agencies (federal, state, and local), media, and foreign organizations with which EPRI has an information exchange agreement. On request, RRC will send a catalog of EPRI reports.

TOPICS

Mechanical maintenance Bolted connections Trouble shooting Leaks Corrosion Fatigue

RESEARCH CATEGORY

Reliability, operations, maintenance, and human factors

Electric Power Research Institute and EPRI are registered service marks of Electric Power Research Institute, Inc.

Copyright © 1987 Electric Power Research Institute, Inc. All rights reserved.

NOTICE.

This report was prepared by the organization(s) named below as an account of work sponsored by the Electric Power Research Institute, Inc. (EPRI). Neither EPRI, members of EPRI, the organization(s) named below, nor any person acting on behalf of any of them: (a) makes any warranty, express or implied, with respect to the use of any Information, appar~tu~, meth~d, or process disclosed in this report or that such use may not Infringe privately owned rights; or (b) assumes any liabilities with respect to the use of, or for damages resulting from the use of, any information, apparatus, method, or process disclosed in this report.

Prepared by Raymond Engineering Inc. Middletown, Connecticut and Looram Engineering Yalesville, Connecticut

f!,D

8 7M

kdJ

87t' !t=r1

e::..fJ

!i!iJ

l! . .lJ

PREFACE

This is a Reference Manual designed to help you solve or prevent bolted joint problems. It is designed for rapid access for use in the field or in the office. The manual tells you how to identify and deal with typical problems like leaks, vibration loosening, fatigue, stress corrosion cracking, etc. It gives you field proven techniques to address these problems.

The manual is NOT intended to be a substitute or alternate for existing bolting specifications, Codes, or standards. We have not knowingly violated any such documents, but if you find that our recommendations are in conflict with an existing Code or standard, you should comply with the latter.

The manual does not, furthermore, cover all possible types of joints or all applications. It describes normal solutions for typical problems. Fortunately, most bolted joints do not cause problems because they are tolerant to wide variations in preload or assembly procedures. This is primarily due to very Conservative design processes. Those few problem joints can benefit from the advice contained in this manual.

A note of caution. The critically important joints where safety and/or performance are of prime consideration should not be addressed solely with the information contained in this manual. The manual provides valuable guidance, but the final judgment must come from a person qualified in the problem area, such as a cognizant engineer, metallurgist, corrosion or fatigue specialist. For example, we don't intend for you to apply the general discussions on gasketed joints to ASME Code Class 1 Joints (see CLASS 1 JOINTS for further details) or for components designed to Division III, Subsection NC Or, the general GASKET STRESS recommendations should not be used if the flanges are made of {'uncommon" materials.

We have tried to define the limitations of each discussion or recommendation contained in the manual, but again, we can't foresee or cover all possibilities. There can be no substitute for good engineering judgment.

Given its highly practical approach, the manual should be irrunediately useful to you. Bolting is largely an empirical art at present, and the experience of others is often your best guide to success. Apply the suggestions contained here thoughtfully and carefully to your own application and you should be able to minimize your bolted joint problems.

I!-.:i-j

8_l""

SYMBOLS AND UNITS

A AB AG Aj

AR As At C D Df E Eb Ej F FAT

FB FG FGA FN

Fp Fpmax Fpmin FTH

HH HN HNI HN2 HT ID K Kb Kj

Accuracy of preload developed at assembly ['Yo] Cross-sectional area of the body of a bolt [in'] Area of gasket [in'] Effective cross-sectional area of that portion of the joint which is loaded by one bolt [in'] Root cross-section area [in'] Tensile stress area [in'] Thread stripping area [in'] Conversion factor [inch-Ibs to ft-Ibsj] Nominal fastener diameter [inch] Distance across flats of bolt head or nut [inch] MODULUS OF ELASTICITY [psi] Modulus of elasticity of bolt material [psi] Modulus of elasticity of joint material [psi] Force [Ibs] Tension in a preloaded bolt as affected by tool accuracy, plus subsequent thermal loads [Ibs] Required assembly bolt load [lbs] Force required to seat the gasket [Ibs] Nominal clamping force on the gasket at assembly [Ibs] The net clamping force on the joint after the system has been pressurized [Ibs] Bolt preload in each bolt at assembly [lbs] Maximum assembly preload [Ibs] Minimum assembly preload [Ibs] Tensile force created in a bolt by differential thermal expansion between bolt and joint [Ibs] Height of the head of a bolt [inch] The thickness of the nut [inch] The thickness (height) of the first nut [inch] The thickness (height) of the second nut [inch] The depth of a tapped hole [inch] Inside diameter of sealing surface (gasket) [inch] Nut factor Stiffness of a bolt [Ibs/inch] Stiffness of a joint [Ibs/inch]

vIII

L LB LC LBG LE LG LP LS LSG LT M n OD P Rm

S Smax STH

SGI SGR SGM Sy Syt T Th Tj To

Cib Cij AJ AL At

Symbols and Units

Nominal length of a bolt [inch] Effective lengh of the body of the fastener [inch] Complete overall length of a bolt [inch] The length of the body within the grip length [inch] Effective length of the fastener [inch] The grip length of the joint [inch] The pressure load on the joint [lbs] The effective length of the threads of a fastener [inch] The length of the thread within the grip length [inch] Length of threaded region of a bolt [inch] Percentage of yield (as a decimal) Number of bolts Outside diameter of sealing surface (gasket) [inch] Internal pressure [psi] Ratio between residual stress on the gasket and the contained pressure Stress [psi] Maximum bolt stress at operating conditions [psi] Stress created in a bolt by differential thermal expansion between bolt and joint [psi] The initial stress on the gasket at assembly [psi] The residual stress on the gasket [psi] Maximum gasket stress at operating conditions [psi] Yield strength [psi] Yield strength of bolts at operating temperature [psi] Torqe [ft-lbs] Operating temperature of the bolt (OF) Operating temperature of the joinWF) Initial assembly temperature of the bolt and/or joint (OF) Coefficient of expansion of bolt material[in/in/oF] Coefficient of expansion of joint material[in/in/oF] Change in thickness of the joint [inch] Change in length of a bolt [inch] Change in temperature (OF)

------------ --"

TABLE OF CONTENTS

Page Number

INTRODUCTION ___ . . . . . . . . . . ... . . . . ... . . . . . . . . . . . . . .. . . . . . . . . 1 ACCURACy................................................... 2

ASSEMBLY PROCEDURES, GENERAL-All Joints............ 7

ASSEMBLY PROCEDURES, GENERAL-Gasketed Joints..... 9

ASSEMBLY PROCEDURES, GENERAL-Non-gasketed Joints. 11

ASSEMBLY PROCEDURES-Qualification ..................... 11

ASSEMBLY PROCEDURES-Stretch Control.................. 14

ASSEMBLY PROCEDURES-Structural Joints................. 16

ASSEMBLY PROCEDURES-Tensioning....................... 17

ASSEMBLY PROCEDURES-Torquing......................... 17

ASSEMBLY PROCEDURES-Turn of Nut .................... · 19

BOLT DIMENSIONS ............................... ····.····.. 22

BOLT MATERIALS ................................... ·· .. ····· 22

BOLTS, BROKEN ........................... ·· .... · .. ·.··· .. ··· 23

CLASS 1 JOINTS ................................. ·· .... ···· .. · 23

COEFFICIENT OF EXPANSION .................... ·· .. ······· 23

CORROSION. . . .. . . . . . . . .. .. . . . . . .. .. . . . . . .. . . . . . . .. . . . . .. . . . . 24

CREEP ......................................................... 26

x Table of Contents

DATUM ROD ........................................... · ..... . 28 DIMENSIONS, BOLT ................................ ·.·· ..... . 27 DISASSEMBLY PROCEDURES ............................... . 27 EFFECTIVE LENGTH ........................................ . 28 ELASTIC INTERACTIONS ............................ ··· .... ·· 28 EMBEDMENT ................................................. . 32 EXTENSO METER ............................................. . 32 FAILURE OF FASTENER ..................................... . 32 FASTENERS, BROKEN ....................................... . 33 FASTENERS, STRIPPED ...................................... . 34 FATIGUE FAILURE ................................... ··· ..... · 34 FINISH, FLANGE ..................................... ·· ..... · 40 FLANGE, CAST IRON ....................................... . 41

FLANGE, OVAL RING ............................ ······ ... ··· 41 FLANGE ROTATION ........................................ . 41 FLANGE, STAINLESS STEEL ................................ . 42 FLANGE-UNCOMMON MATERIALS ....................... . 42

GALLING ................................................... . 43 GASKET, CREEP OF ......................................... . 45 GASKET LEAKS .............................................. . 46

GASKET STRESS ............................. ················· 46 GASKETS, ELASTOMERIC OR PLASTIC. ................... . 50 GASKETS, IN GENERAL. ................................... .. 50 GASKETS, METAL O-RING .................................. . 53 GASKET, METAL (SOLID OR CORRUGATED) .............. . 55 GRIP LENGTH ............................................... . 55 HARDNESS OF FASTENERS ................................ .. 58 HEATERS ........................................ ······· ...... . 59 IDENTIFICATION OF MATERIALS .......................... . 60 INSPECTION OF BOLTED JOINTS ........................... . 81 LEAK RATE ....................................... ···· .... · .. · 62 LEAKS ........................................................ . 64

LENGTH, ACOUSTlC. .......................... ·············· 67 LENGTH, EFFECTIVE. ................................ ·· .... .. 67 LOOSENING ........................................... · ..... . 68 LUBRICANTS .................................... ·· ..... · ..... . 68

MATERIALS, BOLT .............................. ··········· .. 72

MATERIALS, IDENTIFICATION OF ......................... .. 72

MATERIALS, PROPERTIES OF .............................. .. 77

-.-~

~~

~ ~~ , .. ~ :1'~ :1~ ~:±3

:;: :;: :r ti!i t'- ~

t!.-:;;i-j

~~

1lI!i~I'j

IF:~i3

Table of Contents xl

MODULUS OF ELASTICITy.................................. 77 NUT FACTOR................................................. 79 PRELOAD. . .. .. . . . . . . . .. .. . . . . .. . . . . . . . . .. . . . . . . . . . . . . . .. . . . . . 80 PRELOAD, CONTROL OF.................................... 81 PRELOAD, INITIAL........................................... 81 PRELOAD, LOSS OF.......................................... 81 PRELOAD, RESIDUAL ........................................ 84 PRELOAD, SELECTION OF................................... 84 PRESSURE BOUNDARIES......... . . . . . . . . . . . . . . . . ... . . . . . . . .. 94 PREVAILING TORQUE. .. . . . . . . .. . . . . . .. .. . . . . . . .. . . . . . . .. . . .. 95 PROCEDURES, ASSEMBLy/DISASSEMBLy ................... 95 PROOF LOAD........................................... 96 PRYING ....................................................... 96 RELAXATION, FASTENER.................................... 97 ROTATION OF FLANGE...................................... 97 SEALANTS. . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . . 98 STIFFNESS OF FASTENER AND JOINT...................... 99 STRENGTH OF BOLTING MATERIALS ....................... 101 STRENGTH OF FASTENERS-GENERAL ..................... 101 STRENGTH OF FASTENERS-STATIC. ....................... 102 STRENGTH OF THREADS .................................... 104 STRESS AREA ................................................. 105 STRESS CORROSION CRACKING ............................ 105 STRESS RELAXATION ........................................ 107 STRETCH OF FASTENERS.. .. .. . .. .. .. .. .. .. .. .. .. .. .. .. .. ... 107 STRIPPED THREADS .......................................... 113 STUDS, BROKEN .............................................. 113 TEMPERATURE, HIGH ........................................ 113 TENSILE STRENGTH ......................................... 113 TENSILE STRESS AREA ....................................... 114 TENSIONERS .................................................. 114 TENSIONING, HYDRAULIC. ................................. 114 THERMAL EFFECTS ........................................... 117 THERMAL STRESSES ......................................... 122 THREAD STRESS AREAS ..................................... 122 THREAD STRIPPING.......................................... 123 TORQUE, BREAKAWAy ...................................... 127 TORQUE, CONTROL OF ...................................... 127 TORQUE LOSS ................................................ 127

di Table of Cont~rits

rORQUE PROCEDURES ...................................... 127 rORQUE, RE-STARTING ...................................... 128 rORQUE RELAXATION ....................................... 128 rORQUE, SELECTION OF .................................... 128 rORQUE-TURN PROCEDURES ............................ ... 130 rRAINING BOLTING PERSONNEL. .......................... 130 rURN OF NUT .......................................... ··.··· 131 JLTIMATE STRENGTH ....................................... 131 JLTRASONICS ................................................ 131 VIBRATION LOOSENING ..................................... 132 MASHERS, CRUSH ........................ ................... 134 MASHERS, PLAIN ............................... ............. 135 (IELD STRENGTH ............................................ 135

\PPENDIX, GASKET STRESS WORKSHEET .......................... 141 TORQUE COMPUTATION WORKSHEET ... ............. 145 PRELOAD/TORQUE SELECTION WORKSHEET .......... 149 THERMAL STRESS WORKSHEET ............... ......... 153

BOLTING PROCEDURES REFERENCE MANUAL

INTRODUCTION

We've written this manual for people who must disassemble and re-assemble bolted joints in nuclear power plants. It describes bolting practices which should help you identify, understand and then solve or minimize bolted joint problems. We've taken into account the fact that the options available to maintenance personnel are limited. Usuaily you can't re-design equipment to solveproblems. You can't afford to use state-of-the-art bolting tools and instruments on any jOints except those whose past performance tells you "there's no other way to keep them working" -and whose safe performance is critically important.

We recognize that your opportunities to J'experiment" are virtually non-existent. You can't try new techniques on a plant while it's in operation-you'd risk shutting it down. You can't try things during shutdowns-there are far too many things to do. Yet if faced with bolting or bolted joint problems, you must solve them somehow, usually by making small but relatively safe improvements in tooling or procedures during the next shutdown.

This manual describes things you can do to reduce problems and it lists them, wherever possible, by increasing complexity or cost. The

2 Bolting ProL .... uures Reference Manual

basic idea is "try the simple things first. Do the more complex and expensive things later." And just about all suggestions are based on field experience, rather than on theory (though all can be defended and explained theoretically if necessary).

This is NOT a referenCe manual for designers. The theories behind the recommendations are not discussed at any length in the manual. The theoretical literature on bolting is already large and of minimal help to operating and maintenance personnel. This booklet is for people whose prime concern is solving problems, using practical and economical methods and equipment. "Explanations" are given only when needed to clarify the recommendations.

The manual has an "encyclopedia" format designed to make topics easy tolocate. Topics are listed alphabetically and identified by legends that are printed bold. Each topic is described briefly, with typical data if pertinent, and with cross references to related topics (capitalized words in the text-"STRESS RELAXATION", for example-mean that the topic is covered separately.)Thus, if temperature cycles were causing gasketed joints to leak, you'd be directed to "GASKETS" or "TEMPERATURE, HIGH" or "LEAKS" or "THERMAL STRESS" where you'd find a brief explanation of the possible reasons for the problem and recommendations for reducing the problem in the future.

Nor is this a manual of "final solutions" to bolting problems. Bolting is a largely empirical science; a large number of variables are involved in the assembly process. We can never hope to control or even measure the variables for individual bolts-so we must rely heavily on our own past experience, and that of others, to identify and reduce problems. Success is never guaranteed, but things which have worked on similar situations in the past are worth trying and will usually help. Good luck!

ACCURACY

(See PRELOAD, SELECTION OF; PRELOAD, LOSS OF; ELASTIC INTERACTIONS; RELAXATION, FASTENER)

Accuracy is defined as being free from mistake or conforming to a standard. Precision is a term which is closely associated and usually combined with the notion of accuracy. These two terms are best defined by an example:

:t.~ ~

~ ,~

:t:~

~~i~ ~r ~:i"J ~.~

~~ SEi~ ~.:!!

I ~.:3

~i:::!J ~i~ ~i~ ~.~

I ~.~

~::i!I I

~i~ ~i~ ~~ @::r~

\---Bolting Procedures Reference LVl~nual 3

The torque output of a wrench is measured on a calibrator five times in succession at rated torque. The results of the exercise are described as follows:

o The average output of the wrench is found to be 5% low. The wrench is said to be 5% inaccurate.

o The output from test to test varies over a range of + / -1 %. The wrench is said to have a precision of 1%.

o The performance of the wrench may be characterized by stating that the expected torque output has an accuracy of 5% and a repeatability of 1%.

Sources Of Error

The preload developed in a fastener at assembly is affected by many factors:

Tool Accuracy The accuracy of the tool output (i.e., the accuracy of the torque wrench in the example).

Operator Operator accuracy relates to the error introduced by the operator as a result of such things as skill, working conditions, hardto-reach nut locations, lighting and carelessness.

Control Control accuracy relates to the accuracy with which the preload is controlled by the quantity actually being measured. For example, when torque is used to control preload, the control accuracy is determined by the torque/preload relationship (see PRELOAD, SELECTION OF).

Relaxation During the jOint assembly and throughout the service life of the joint, the fastener preload changes. The joint make-up procedure has a major influence on the fastener preload (see PRELOAD, LOSS OF).

Typical Preload Accuracy Obtained with Various Tools

Table A shows typical preload accuracies for various tightening methods. In general, the accuracy depends on the joint configuration, the type of tool, and the make-up procedure employed. Accuracies cited are for initial preload in individual bolts (See PRELOAD, INITIAL).

4 Bolting Procedures Reference Manual

TABLE A Preload Accuracies

Tightening Procedure

Torque

Torque

Impact Wrench

Hydraulic Tensioner

Hydraulic Tensioner

Stretch Control

Turn of Nut

Operator Feel

Comment

Calibrated torque wrench, lubricant and torquing procedure calibrated by stretch or load measurements on the actual joint.

Hard Joint: Metal-to-metal contact between 2 or 3 plies of joint components

Soft Joint: Gasketed or hard joint with many plies.

Calibrated torque wrench. Target torque value determined from friction estimate from literature.

Hard Joint Soft Joint

Impact wrench, air driven or slugging wrench or hammer.


Procedure calibrated by using stretch measurements on the actual joint


Tensioner pressure set at a value equivalent to the target preload. Tensioner inefficiency and relaxation not considered.


Datum rods used in heater holes.

Micrometer depends on the length of the fastener and the end conditions

Ultrasonic stretch measurements

Structural steel joints using ductile bolts (A325 or A490). Joint fitted carefully. High snug torque followed by measured turn of nut.

For bolts less than 1" D.

Preload Accuracy (%)

+/- 17 to 23%

+/- 20 to 40%

+/- 20 to 40% +/- 30 to 70%

+/- 45 to 800Al +/- 60 to 100%

+/- 10 to 20% +/- 20 to 40%

+/- 30 to 50% +/- 40 to 700/0

+/- 101020%

+/- 51030%

+/- 21025%

+/- 10 to 20%

+/- 100 to 200%

~~,

-----------------------------Bolting Procedures Reference ~.~,£~ual 5

Figures 1 and 2 show the joint classifications, either hard or soft.

FIGURE 1 Typical "hard" joints. Note that metal-to-metal contact Is Involved In each.

FIGURE 2 Typical "soft" jOints. The one on the left Is soft because It involves many plies andl or overSized or slotted holes.


Accuracy Specifications

What is the preload accuracy required for successful joint perfonnance? Thanks to the complex behavior of bolted joints, a universal answer

to this question is impossible. It is felt that both the average and variability of preload across a bolted joint affect the performance of the joint.

Here are rough estimates of typical accuracies often specified for various types of joints:

JOINT TYPE

Casketed joints Large diameter (grea ter than 24") Small diameter (less than 24")

Hard joints

TYPICAL ACCURACY SPECIFICATION

+/- 10 to 30%

+/- 20 to 40%

+/- 15 to 30%

NOTE: These accuracy specifications for initial preloads in individual bolts are from experience and are highly dependent on the design of the joint and the service conditions. Residual preloads will often vary a lot more than this (See ELASTIC INTERACTIONS).

To Improve Accuracy

o Qualify the joint make-up procedure by measuring the preload achieved in the actual joint.

o Choose tools and procedures from Table A which are more accurate than the present methods.

o Preload results can be improved by controlling the variables which affect the control element; i.e., friction, surface finish, etc. (see assembly procedures for the specific tightening method of interest).

References The following documents and texts have helped us prepare this section-and can give you additional information.

1. Bickford, j. H., Section 23-Bolted and Riveted Joints. Standard Handbook of Machine DeSign, Editors j.E. Shigley and C.R. Mischke, McCraw Hill, New York. 1986

2. Yahr, G.T. Preloading of Bolted Connections in Nuclear Reactor Component Supports. NUREG/CR-3853, ORN-6093. Nuclear Regulatory Commission. 1984.

Bolting Procedures Reference Manual 7

ASSEMBLY PROCEDURES-GeneralAll Joints

The designer can do just so much to determine the behavior and operating life of a bolted joint. The rest is up to the people who assemble it. If the joint is not "properly" assembled, it won't perfonn as intended. Unfortunately, the assemblers must cope with a large number of variables which neither they nor the deSigner can ever hope to controlvariables which affect the outcome. They include the smoothness, hardness, lubricity of all surfaces, the condition of the parts (rust, tool marks, defects ), the calibration of the tools, the accessibility of the bolts, the environment in which the mechanics must operate and many other things.

Specific suggestions are given below for various types of assembly procedures using such control means as torque, tension, bolt stretch. Regardless of which method you use, you should also take the steps listed here if you need and want to optimize results.

There's no magic bolting procedure that avoids all of the variables which influence the results. If your present procedures have worked, there's no need to change them. If they haven't worked in some places, however, the procedures described below and in the following sections on specific types of assembly may help you improve your own procedures and results.

Basic Recommendations

Be Consistent! There are enough "uncontrolled" variables involved as it is; don't make things worse by adding to them. Wherever possible, use the same tools in the same manner.

Train your bolting crews! Show them why good practices make a difference. Warn them of the problems you'll encounter if things are not done properly. Experience shows that training is more apt to improve bolting results than is the use of more expensive tools or procedures (see TRAINING OF BOLTING PERSONNEL).

Supervise the work, especially on critical joints! This is almost as useful as good training in optimizing results.

Keep your tools in good shape! Rebuild them periodically. It is counterproductive and demoralizing to stop a job for tool repairs.

Use written procedures! The procedures should include:

o joint identification by identifying number, system, location, material, size.


D Fastener identification including size and grade. D Tool identification and verification of current calibration

sticker. D Detail all assembly steps:

-Cleaning of parts; solvents, rags, brushes -Lubrication-type and grade, application -Visual inspection of components for damage -Give tool settings; i.e., pressure, torque or turn -Specify the tightening sequence along with the tool

setting for each pass. -Provide sign-off spaces for the crew performing the work,

supervision and Q. c.

Preparation of Bolting Materials

Wire-brush studs and nuts (when needed) to remove any dirt on the threads. Use stainless steel bristles on alloy materials.

Visually examine studs and nuts after cleaning to assure freedom from burrs. Nuts should turn freely on the studs a distance equal to their in-service make-up. If any burrs are present, perform One of the following steps:

D File off burrs of a minor nature. Files utilized for alloy materials should not have preViously been used on carbon steel materials.

D Chase threads with a tap and die. D Return the nut and/or stud to the store room and draw a

new one of the same size, type and qualification for installation.

Upon completion of the cleaning operations, coat studs with a film of an approved lubricant prior to installation.

Coat the bearing surface of the turned element (the nut or the bolt head) with the approved lubricant.

It is always desirable to use hardened washers between the turned element and the joint surface. In some codes this is a requirement; and a hardened washer, under the element not turned, may also be required.

" 1--

Bolting Procedures Reference Man ual 9

ASSEMBLY PROCEDURES-GeneralGasketed Joints

(See ASSEMBLY PROCEDURES-General-All Joints)

Preparation of Facing Finish

Clean gasket seating surface using suitable solvent and wire bristle brush (use stainless steel bristles on alloy components).

After cleaning, visually inspect the seating surface for defects such as radial scores.

Inspect the seating surface for warping. The two illustrations below show flange conditions that can cause

a gasketed joint to leak. Don't exceed the conditions shown.

FIGURE 3 The conditions shown can result in leak. ,,1 + ,,2 should not exceed 0.015" for best results.

OTHER PROBLEM AREAS


FIGURE 4 For metal-jacketed or spiral-wound gaskets the maximum total deviation shown above should not exceed 0.015". Compressed asbestos or rubber gaskets can stand deviations as high as 0.030". Solid metal gaskets, on the other hand cannot be used safely if the deviation exceeds 0.005".

Flange Alignment and Gasket Installation

o Once flanges have been lined up, visually examine them to assure that an acceptable fit has been obtained.

o Install a few studs in the flanges to maintain alignment of the flanges. Leave room for insertion of the gasket.

[J Visually examine gasket prior to installation to assure it is free of defects. Return defective gaskets to the store room and draw a new one of the same size, type and qualification.

o Carefully insert gasket between the flanges to assure proper placement and prevent damage to the gasket surfaces.

o Install remaining studs and screw nuts on hand-tight. o For the rest of the assembly procedure, see the section for

the tightening method being used (for example, Assembly Proced ure-Torquing).

-'~'>'

\ Bolting Procedures Reference M J 11

ASSEMBLY PROCEDURES, GENERAL-Non-gasketed Joints

(See ASSEMBLY PROCEDURES-GENERAL-All Joints)

ASSEMBLY PROCEDURES, QUALIFICATION

Bolted joint assembly methods using torque, turn-of-nut, or hydraulic tensioners result in fastener preloads which are uncertain (see ACCURACY and ELASTIC INTERACTIONS). During the service life of a joint the original fastener preload can be significantly altered by external loads and service conditions (see THERMAL STRESS, STRESS RELAXATION, FLANGE ROTATION, GASKET CREEP).

If a bolted joint is to be successful, preloads developed at assembly must be sufficient to maintain joint integrity throughout the service life.

An assembly procedure is qualified if the assembly method reliably produces fastener preloads which result in successful joint performance. This qualification should encompass the component parts of the joint (nuts, bolts, gaskets, washers), tools, personnel and the method of making up the joint.

There are at least three ways to qualify an assembly procedure; by experience, proof test, or by experimentation.

Experience

The vast majority of bolted jOint assemblies are assembled without formal documented procedures. The assembly method is left to the mechanic's judgement. For many joints this assembly practice has been satisfactory; the jOints have been successful in service.

If a joint has been assembled by a procedure or by a mechanic's feel, and the operational experience has been good, then the assembly procedure can be said to be qualified by experience or precedent.

12 Bolting PI,- "lUres Reference Manual

Proof Test

The most common method of qualifying a joint and the assembly method is a proof test. Normally a load or pressure of 1.25 to 2.0 times the design load is applied to the assembled jOint. The philosophy of the test is that if the joint successfully carries the proof load, then it will be successful in service. Since the joint is assembled prior to the test the assembly procedure is also qualified by the test.

Experimentation

Some assembly procedures are qualified by extensive experimental programs. The TIJRN OF NUT assembly method for assembling structural steel using A325 or A490 bolts is an example of an experimentally qualified procedure. The procedure was qualified by assembling many joints and loading them to failure.

These three qualification approaches work well for the vast majority of joints. There are, however, a small percentage of joints which will pass a proof test and then fail in service. These failures indicate a deficiency in the joint design, the assembly procedure, or both. Certainly the joint is not qualified for service, and the qualification method is suspect.

When faced with a joint which fails in service, a more detailed approach to qualifying the assembly method may be beneficial. The following is an outline of points which must be addressed.

Joint Assembly Qualification-Which Joints

o Joints having a history of service failures should be qualified.

o Critical joints. Class I pressure boundary and safety related joints.

Design Requirements

Review the joint design requirement using Level 4 and Beyond Level 4 sections of PRELOAD, SELECTION OF as a guide. Some of the areas to be reviewed are:

rJ Materials: Specification, type, grade, hardness

\---Bolting Procedures Reference l\.~ . .. J.al 13

o Critical Dimensions: Diameters, thickness, parallelism, surface finish

o Preload Requirements: Torque, turn, tension, stretch, gasket deflection

o What is the design preload? The maximum and minimum acceptable preload? The acceptable average and scatter? What provisions, if any, have been made to accommodate or compensate for changes in load due to service conditions such as THERMAL STRESS, CREEP, FLANGE ROTATION, STRESS RELAXATION.

o The design preload requirement should consider the assembly, the load at operating pressure and temperature, and the loads remaining at the end of service life.

Operating Experience-Assembly

Do procedures exist? Do the procedures contain basic elements of good practice? Will the procedures get the preloads close to the deSIgn requirements? (See ASSEMBLY, and ACCURACY.)

Are the procedures followed? Don't just review the paperwork, go out and witness the disassembly and assembly.

Operating Experience-Disassembly

Are there disassembly procedures? (See DISASSEMBLY) What is the state of load in the fasteners? This may be measured by stretch measurements or by breakaway torque as the fasteners are unloaded (see INSPECTION OF BOLTED JOINTS).

What is the condition of the joint components? Gasket adequately compressed? Buckled? Sealing surface cut, corroded, warped?

If the outlined review fails to identify any glaring deficiencies, the next step is to verify that the design preloads are achieved at assembly. The preloads may be measured by using an extensometer to measure the fastener stretch (see STRETCH, STRETCH MEASUREMENT OF, EXTENSOMETER), or some other load measuring device.

If the developed preloads are not within specification, the assembly procedure must be modified to produce more reliable results (see ACCURACY, ASSEMBLY PROCEDURE-GENERAL).


ASSEMBLY PROCEDURES-Stretch Control

(See ASSEMBLY PROCEDURES, GENERAL-All joints; STRETCH OF FASTENERS)

General

Monitor andlor control anything which affects the stiffness of the bolt since this determines the relationship between stretch and preload (See STIFFNESS OF FASTENER AND JOINT). For example, try to maintain uniformity of:

o GRIP LENGTH of the assembly. o Body length and thread length of the bolts (often varies

from supplier to supplier). o MODULUS OF ELASTICITY of bolts (often varies from lot

to lot).

When making the stretch measurements consider the following:

o The temperature of the fasteners; affects both length and (if you're using ULTRASONICS) the velocity of sound.

o If using C-mics, take several readings in different quadrants and average the results, or measure over small balls pressed into the center of both ends of the fastener.

o If using depth mics, always orient them the same way. Ensure that the seating configuration for the DATUM ROD is clean and reliable.

o If using an ultrasonic EXTENSOMETER to measure stretch, read and follow the manufacturer's instructions.

Specific Procedures

Prepare the fastener ends to improve the accuracy of the stretch measurements. The stretch measured will be on the order of thousandths of an inch (see STRETCH OF FASTENERS).

CJ Accuracy will be improved if you face the ends flat and parallel to each other with a surface finish of 125 microinches.

[J For depth rod measurements, ensure that the DATUM ............. r-.. ___ L !n !_,-,,, nt .4o:>hriq.


Reco~d the desired stretch and the recommended tool setting (torque or tenSlQner pressure).

o If the preload requirement is given as a load lIb], convert it to a stretch (see STRETCH OF FASTENERS).

o If the requirement is in torque, contact the vendor for the load specification, which may then be converted to a stretch (or see PRELOAD, SELECTION OF).

Measure and record the initial length of each fastener. Also measure the temperature of each fastener or a representative sample of the fasteners in the joint.

o Check the repeatability of the stretch measurement by making repeated length measurements on a few studs. The length measurement should be repeatable within 1.0 to 2.0% of the desired stretch.

For joint make-up, follow procedures similar to those in ASSEMBLY PROCEDURES, GENERAL-Gasketed joints or ASSEMBLY PROCEDURES, GENERAL-Structural joints.

After the third pass, measure the stretch of the last four fasteners tightened. Average these four stretches (JlL3).

Calculate .the torque or tensioner pressure (T4) for the next pass as follows:

T4 ~ T3 x JlL4/JlL3 Where: T4 ~ Torque or tensioner pressure for Pass 4

T3 ~ Torque or tensioner pressure for Pass 3 Jl L4 ~ Desired final stretch JlL3 ~ Average stretch of the last four fasteners tightened during Pass 3

Put the fourth pass on the joint using the tool setting determined above.

After the fourth pass, measure the stretch of all fasteners. Adjust the studs as necessary.

o Bring the low loads up by re-tightening at tool setting T4 or by increasing T4. Remember that as you increase the load on the low ones, the higher ones will lose load due to elastic interactions (see PRELOAD, LOSS OF). If the high ones are marginally in specification, it is wise to apply a slight increase in stretch to these to compensate for the expected interaction loss as the low ones are torqued.

o C~n~inue this p.rocess until the stretch values converge wIthm the specified tolerance .


NOTE: Measure the fastener temperature each time a stretch measurement is made. If the temperature changes from the initial measurement, a correction may be required (see STRETCH OF FASTENERS).

ASSEMBLY PROCEDURES-Structural Joints

Line up joint members using drift pins in a few holes. Install bolts in the remaining holes hand-tight. Then snug them all,

including the first few installed. Snugging torques should be 10-30% of final torque and should be applied first to the bolts in the most rigid part of the pattern, then to those farther out, then to those on the free edges of the joint.

Check to see that the joint members are fully pulled together. If not, apply more snugging torque.

If drift pins have been used and the joint is not under load, knock them out at this time. Install and snug remaining bolts.

Apply final torque to each bolt, again starting at the most rigid part of the joint and working towards the free edges.

If drift pins have been used in a joint under load, knock them out after all other bolts have been fully tightened. Then install and fully tighten the remaining bolts.

If you're especially concerned about the performance or safety of this joint, it would be wise at this point to re-apply final torque to all of the bolts in the joint, following a reverse sequence (starting at the edges and working towards the most rigid paint), or use some other preload measuring device (STRETCH OF FASTENERS).

Bolting- Procedures Reference Manual 17

ASSEMBLY PROCEDURES-Tensioning

(See ASSEMBLY PROCEDURES, GENERAL-All Joints) Follow the tensioner manufacturer's instructions and these general

recommendations:

o Use multiple tensioners ganged together, if possible to reduce ELASTIC INTERACTIONS.

o Use uniform run-down torque. Be sure that the nut turns. o Verify that the specified hydraulic pressure is applied to

the tensioner. -Check all hydraulic connections. -When using multiple tensioners ganged to one pump,

running one nut down may reduce the system pressure. Check the pressure before running the nut down on each stud.

-The nominal tensioner load should be 20 to 30% higher than the desired stud preload. This over-tension compensates for the relaxation effects (see TENSIONERS).

-The make-up procedure should call for a final check pass after the specified tightening passes. The check is accomplished by applying the final tensioner pressure to each stud and attempting to run the nut down. If the nut moves, this indicates that the residual preload was low and more tightening passes are reqUired.

ASSEMBLY PROCEDURES-Torquing

(See ASSEMBLY PROCEDURES, GENERAL-All Joints)

General

Ensure that the load bearing surfaces are in good condition. Check the thread flanks, bearing surface of nuts or bolt head, washers and the flange surface.

Gean and lubricate threads and nut and/or,bolt head bearing surface. Use specified lubricant and apply it uniformly as directed.

•


Using hardened washers under the turned element is recommended and required by some codes. This improves the torque/preload relationship.

Always run the nuts or bolts down by hand. If they won't run by hand, there is a defect in the thread.

Use a multiple pass, cross-bolting procedure (see below). If there's a gasket, ensure that it is compressed evenly. Use caliper

measurements in four quadrants if possible. Torque as many studs as possible simultaneously. This will help pull

the joint down evenly and reduce loss of bolt load during the pass (see PRELOAD, LOSS OF).

All torque wrenches should be of adequate capacity and recently calibrated.

All thread LUBRICANTS and anti-seize compounds should be of an approved type. It's useful to limit the types of lube used in the plant. This helps avoid mistakes and lets you build experience about the characteristics of the selected lube.

Apply torque at a uniform rate; final torque should be reached "in motion". If you exceed the final torque, loosen the fastener a little and try again to hit the final point While in motion.

Hold torque wrenches perpendicular to the axis of the bolt while torque is being applied.

If hydraulically powered torque wrenches are used, ensure that adequate reaction points are provided.

Modified Tightening Sequence

D Better preload uniformity can sometimes be achieved by torquing the fasteners in the reverse sequence in the final pass. If you have had previous troubles with this joint, try this after the prescribed passes have been accomplished.

In critical situations you can verify the preload achieved by making stretch measurements of the fastener.

Speciflc-Gasketed Joints

Torque bolts or nuts in a cross-bolting pattern. Torque the joint using a minimum of four torquing passes, using

a cross-bolting sequence for each pass. The torque values for each sequence are given. below:


Pass 1 Bring all nuts up hand-tight; then snug-tight evenly. Pass 2 Torque to a maximum of 30% of the final torque (See

PRELOAD, SELECTION OF). Check that the flange is bearing uniformly on the gasket.

Pass 3 Torque to a maximum of 60% of the final torque. Pass 4 Torque to the final torque.

After the four basic torquing passes are completed, continue torqUing the nuts using the final torque in a clockwise manner until no further rotation of the nut is observed. This process may require an additional five to seven passes.


1. Bolt TorqUing. Diablo Canyon Power Plant. PG&E. Number MPM-54.1, Rev. 1. July 31, 1984.

ASSEMBLY PROCEDURES-Turn of Nut

(See ASSEMBLY PROCEDURES-General)

General

Make sure that the bolts being tightened have sufficient ductility to be tightened past yield. ASTM A325 and A490 materials are examples of acceptable materials.

Ensure that the joint components are correctly assembled. Things to look for:

Bolt: Proper type and length Washers: Standard, bevel or heavy. Ensure that the

correct washer is installed where required. Joint Alignment: Use drift pins to line up the joint members.

Mark the socket or turned element of the fastener (nut or bolt head) so that turn may be measured.


For best results, apply a high snugging torque; then measure turn. This torque should produce 10% to 30% of the minimum specified preload (see TORQUE, SELECTION OF).

Specific Procedures

When the turn-of-nut method is used to provide the bolt tension, first bring enough bolts to a "snug-tight" condition to ensure that the parts of the joint are brought into good contact with each other. "Snug-tight" is defined by the AISC as the tightness attained by a few impacts of an impact wrench or the full effort of a man using an ordinary spud wrench; but you'll need more than this if the bolts are much over 7/8" in diameter.

Following the initial snugging, place bolts in any remaining holes in the connection and snug them. Tighten all bolts in the joint additionally by the applicable amount specified in Table B progressing systematically from the most rigid part of the joint to its free edges. During this operation, there must be no rotation of the part not turned by the wrench.

SA 307, Grade A bolts should be tightened snug tight only without additional nut rotation.

TABLE B Nut Rotation

Effective Bolt Disposition of Outer Faces Length of Bolted Parts

One Face Normal Both Faces (Distance From to Bolt Axis and Sloped Not More Inside Face of Other Face Than 1 :20 From Bolt Head to Sloped Not More Normal to Bolt Outside Face of Both Faces Than 1 :20 (Bevel Axis (Bevel Nut Plus One Normal to Bolt Washer Not Washers Not Thread) Axis Used) Used)

Up to and in· 1/3 turn 1/2 turn 2/3 turn eluding 4 dia· +/-30 deg. +/-30 deg. +/-45 deg. meters

Over 4 diameters 1/2 turn 2f3 turn S/6 turn but not exceeding +/-30 deg. +/-45 deg. +/-45 deg. S diameters

Over 8 diameters 2/3 turn 5/6 turn 1 turn but not exceeding +/-45 deg. +/-45 deg. +/-45 deg. 12 diameters

~'. !._-

Bolting Procedures Reference Ma ..... al 21

The Table of Nut Rotation is applicable for ASTM A325, ASTM A490 bolt materials. These bolts are designed to be installed to a minimum preload which is 70% of tensile strength. Turn-of-nut procedure using the Table will reliably produce these loads. The bolts will most likely be loaded above the proof strength. Some bolt breakage may occur at assembly. Bolt breakage is acceptable. Simply replace the broken bolt.

Train personnel performing this activity to perform the tightening as directed.

For threaded fasteners requiring locking devices, elastic stop nuts (when compatible with service temperature), lock nuts, jam nuts and drilled and wired nuts are all acceptable locking devices. Upset threads (by peening or other approved methods) may also serve as locking devices. .

Nut rotation is relative to bolt, regardless of the element (nut or bolt) being turned. For bolts installed by 1/2 turn and less, the tolerance should be + /- 30 degrees; for bolts installed by'/, turn or more, the tolerance should be + /- 45 degrees.

To establish a turn-of-nut procedure for bolt length exceeding 12 diameters and/or bolt diameters of greater than 1 1/2 inch, or bolts of materials other than ASTM A325 or ASTM A490, the required rotation from snug is determined by test. Use a suitable tension measuring device in a joint which simulates the actual joint conditions.


1. Specification for Structural Joints Using A325 and A490 Bolts. Research Council on Structural Connections. Pub!. by the AISC. 1980.

2. High Strength Bolting for Structural Steel Joints. Bethlehem Steel Corporation, Bethlehem, Pennsylvania, 1980.

22 Bolting Pl ....... :edures Reference Manual

BOLT DIMENSIONS

The Figure below identifies the commonly used dimensions of an hex head bolt.

FIGURE 5 Symbols used to define various parts of a standard hex head bolt.

< ,:~~ /f-LSG-l

11111111l! D

) HH r- LBG+- LT T D = nominal diameter L "" nominal length

LBG "" length of body LC = overall length LG "" grip length

HOLT MATERIALS

LC

L T = length of threads LSG '" length of threads

within grip HH = height of head HN = height of nut

(See MATERIALS, IDENTIFICATION OF; MATERIALS. PROPERTIES OF)

1--Bolting Procedures Reference rv~.. ..Jar 23

BOLTS,BROKEN

(See FASTENERS, BROKEN)

CLASS 1 JOINTS

A "Class 1" joint is a bolted joint in any item designed and built in accordance with the rules of Subsection NB, Division 1, Section III of the ASME Boiler and Pressure Vessel Code. These rules deal with items whose failure would violate the primary pressure boundary. Because of their importance, these joints deserve the extra care specified in the ASME Code. The "cookbook" procedures of this reference manual should not be applied to Class 1 joints unless good engineering judgement shows that the advice is applicable for the specifiC application.

COEFFICIENT OF EXPANSION

If a bolt gets hot it'll expand-get longer. Figure 6 shows the coefficient of expansion in inches per inch per degree Fahrenheit for several common bolting materials. See THERMAL STRESSES for procedures for calculating the effect of this expansion on bolt stresses.


FIGURE 6 Coefficient of thermal expansion versus temperature (all values are approximate).

"

'00 '00 '00

TEMPERATURE 'f

A = A193 816, A540 821 B ~ A193 B7, A320 L7, L7M, L7B, L43 G~A193B5

D~A193B6

CORROSION

G ----,

'00 .00

E ~ A193 B8, B8A F ~ A193 B8T G ~ A193 B8G H ~ A193 B8R, B8RA

Corrosion is a very complex subject; there are many different ways in which an environment can attack and degrade a fastener andlor joint members. There are many types of corrosion-such as erosioncorrosion, fretting, crevice corrosion, pitting, galvanic corrosion. It often takes an expert to identify the type, cause and cure in a specific situation,

Only two types of corrosion are known to have caused problems in nuclear plants: STRESS CORROSION CRACKING and general corrosion of carbon and alloy steel fasteners in the reactor coolant pressure boundary caused by a borated water or boric acid environment.

fiii...,..c:!:":"]

~=:,;;£j

~

~"'~

@:.''l

~---~,

------------------------------ :--Bolting Procedures Reference Manual 25

Borated Water Corrosion

Borated water or boric acid corrosion can best be described as accelerated general corrosion.

Only a small amount of laboratory data on boric acid corrosion exists and much of the relevant data were only recently published. In the concentration used in reactor coolant systems, boric acid is a relatively weak add. However, under wetting and drying conditions, boric acid may concentrate in a slurry forming a saturated solution. The available data indicate that corrosion rates as high as 1.7 inches per year (reduction in diameter of cylindrical specimens) may result when carbon and alloy steels are exposed to borated environments under these conditions. Corrosion rates are rapid at 200 degrees F and a Ph of approximately 3. The corrosion process may be active at service temperatures (greater than 350 degrees F) because of localized cooling of hot fasteners.

There do not appear to be any differences in corrosion rates for the common carbon and low alloy steel bolting materials. Coatings, platings, and various surface treatments have generally not provided adequate corrosion resistance. Corrosion resistant materials such as austenitic and martensitic stainless steels and high strength nickel-base alloys offer good resistence to boric acid corrosion. These materials are now used in many fastener applications, but concerns about strength, degradation in toughness or other forms of corrosion have precluded their general use, particularly in the larger sizes. It should be noted that without the leaking coolant, low alloy steel fasteners have demonstrated exemplary performance.

Solving Corrosion Problems

Experience suggests that the simplest way to prevent corrosion is to prevent leaks. Preventing leaks is not necessarily easy, but it's more economical than replacing materials.

Perhaps the next best thing to do is to let them corrode a little and periodically replace them with new bolts or studs of the same type. If that's not sufficient, replace the fasteners with ones made of corrosion resistent material. An analysis of the component will be required before changing materials.

In general, corrosion problems can't be. solved by amateurs. See an expert!



1. Merrick, E. A. Rivers, J. Bickford and T. Marston. Prevention of Bolting Degradation and Failure in Pressure Boundary and Support Applications. Paper given at SMIRT-81post SMIRT Conference, September 1985.

CREEP

(See GASKET, CREEP OF)

DATUM ROD

A device used to measure the STRETCH OF THE BOLT. A rod is placed in a hole drilled along the axis of the bolt. The far end of the rod is fastened to the bolt; the rest is free. As the bolt is tightened it stretches up around the rod. By measuring the change in the distance between the end of the bolt and the end of the rod, you can tell how much the bolt has stretched and can, therefore, estimate bolt preload. A depth mic is used for the measurement.

Bolting Procedures Reference Mal. ___ .i 27

FIGURE 7 A depth micrometer and datum rod can be used to measure the stretch of a bolt as shown In the illustration.

DIMENSIONS, BOLT

(See BOLT DIMENSIONS)

DISASSEMBLY PROCEDURES Most plants have extensive, written procedures for the assembly of

bolted joints, but few pay much attention to the techniques used to disassemble them. This is usually acceptable, but there are some points you should keep in mind.

There is an enormous amount of elastic energy stored in a bolted flange. The joint, in effect, is a giant spring held compressed by the bolts. As you start to remove bolts, the stored energy of the joint loads the bolts still remaining in the joint. The joint can be distorted, and gaskets or flange surfaces can be damaged. In a lew cases, the final bolts remaining in a £lange have actually been broken as they attempted to hold the expanding joint together.


If you think that you've experienced problems caused by too casual disassembly procedures, you might consider taking them apart with a reverse of the procedure you used to tighten them. Use several passes. Partially loosen each bolt before further loosening any of them. Use a cross-unbolting procedure. Handle the bolts and flange members carefully to avoid damaging them. These steps could save you problems when you come to reassemble the joint.

EFFECTIVE LENGTH

(See LENGTH, EFFECTIVE)

ELASTIC INTERACTIONS

(See PRELOAD, LOSS OF) Tightening one fastener will often partially relieve (in effect, loosen)

previously tightened fasteners near it. Such "cross talk" between fasteners, during assembly or disassembly, is called Elastic Interaction.

Elastic interactions are one of the most common and most extreme forms of PRELOAD LOSS, especially in gasketed joints. They can easily result in a 5:1 or 10:1-or worse -scatter in the residual preload in the bolts in a flanged joint, as shown in the figures below.


FIGURE 8 The "x" marks show the preloads achieved In the Individual bolts In a raised face flange when first tightened with a uniform torque of 275 It-Ibs. The numbers indicate the order In which the bolts were tightened and their relative position on the flange. Bolt #2, for example, Is 180 degrees away from bolt 111. The "0" marks show the residual preload remaining In each bolt alter all of the bolts in the Joint had been tightened (when we were ready to start the second pass). Elastic Interactions between bolts had substantially reduced the preload In many of them.

11 ,. 16 9 II I I) 10

lOLl !'OsmON ~NO NUM'U

FIGURE 9 The Initial (x) and residual (0) preloads In the bolt of our raised face flange following a second pass at 550 It-Ibs. As in Figure 8, the initial bolt-by-bolt preloads are fairly conSistent, but the residual preloads vary a great deal because of elastic Interactions.

o

9

x ~ , 0

\/ 12 5''''3 16 • '1 6 13 15 8 10

BOLT POSITON AND NUMBER


FIGURE 10 Initial (x) and residual (0) preloads In the Joint after a third pass at 825 ft-Ibs.

Cl

" • 9

v"J\/~-\I\// 80

60

40

20,0

12 14 16 , " 6 13

BOLT POSITION AND NUMBER

FIGURE 11 Resldu~1 preload after a final pass at 875 ft-Ibs; at end of Job (x) and after 12 hours (0).

"'0 ;( !! c "

15 B

:: rf'v'{/y'-\j\/~' ~ 8

20.0

5 14 3 9 " 4 B

BOLT POSITION AND NUMBER

Many people find this hard to believe. If 5:1 scatter in preload is common, why don't more joints leak? The answer is that gasketed jOints have a "memory"; their leak behavior depends as much if not more on the initial seating stress on a gasket than it does on residual stress (see GASKET STRESS). The leak behavior is also strongly related to the flange stiffness.


Compensating for Elastic Interactions

In most cases it's not necessary to compensate for elastic interactions. In a few cases, where flanges are relatively flexible, or where there are serious THERMAL EFFECT problems, it will be helpful to reduce the scatter in residual preload. There are a couple of things you can do.

Retighten the Joint Repeatedly

There is no magic torquing or tensioning procedure which eliminates elastic interactions; however, making a large number of passes at the final torque or tension can reduced scatter. Some people keep retightening them "until they no longer move". Doing one or more final passes in a different sequence-for example, the last ones first-can also help.

Measure and Control Residual Preload

The only infallible way to compensate for interactions is to measure and control the residual-rather than just the initial-preload in the bolts. At the present time this means some kind of STRETCH CONTROL: DATUM RODS, for example, or strain gages or load cells.

ULTRASONICS can also be used. If you measure residuals, you'll find that you can achieve relatively uniform residual preloads in the bolts in a joint (+ 110-20% for example) only by applying a different amount of torque to each bolt.

References The follOWing documents and texts have helped us prepare this section-and can give you additional information.

1. Bickford, J. H., Ultrasonic Control of Bolt Preload. 1981 Proceedings-Refining Department. 46th Mid-year Meeting of the American Petroleum Institute. Chicago, Illinois, May 14, 1981.

2. Bickford, J. H., That Initial Preload-What Happens To It? Mechanical Engineering, ASME, New York, October 1983.


EMBEDMENT

(See PRELOAD, LOSS OF) Fastener and joint surfaces are microscopically rough. When first as

sembled, these surfaces (nut, bolt threads, joint members, etc.) only contact each other on high spots. This is insufficient area to support the high loads created by tightening the fasteners, so the high spots will creep and flow. The fastener and jOint members settle in together. The process is called embedment or embedment relaxation, and can cause loss of initial preload of 2-10%. It's worse if the parts are previously unused and/or if the number of jOint plies is large, or tensioners are used to tighten the joint (threads are partially embedded during torquing but not during tensioning).

EXTENSOMETER

A device used to measure the change in length of a test specimen or a bolt. Mechanical, electro-magnetic and ultrasonic instruments are available.


1. Bickford, J. H. Ultrasonic Control of Bolt Preload. 1981 Proceedings-Refining Department. 46th Mid-year Meeting of the American Petroleum Institute, Chicago, illinois, May 14, 1981.

FAILURE OF FASTENER

(See FASTENERS, BROKEN; FATIGUE FAILURE; VIBRATION LOOSENING; STRESS CORROSION CRACKING; THREAD STRIPPING)


FASTENERS, BROKEN (See THREAD STRIPPING)

Causes

Fasteners break when the forces on them exceed their strength. The forces, and the type of "strength" involved, can be static or slowly changing, as discussed in STRENGTH OF fASTENERS-Static. Or the forces can be dynamic, and we must be concerned about fatigue strength, as discussed under FATIGUE FAILURE. Sometimes chemical attack weakens the fasteners and hastens their failure. In most such situations the corrosion will be visible (see CORROSION). In more difficult-to-detect situations, the failure is by STRESS CORROSION CRACKING, caused by a combination of load and corrosion at the tip of a crack.

Prevention

If you're troubled by broken fasteners, you should:

D Determine the basic cause of failure. If pertinent, refer to the sections on CORROSION or STRESS CORROSION CRACKING or FATIGUE FAILURE for suggested steps to reduce the problems.

o If the failures occur under static or slowly changing loads (as the fasteners are tightened, for example), consider taking the following steps, listed in order of increasing complexity or cost:

Reconsider the torque used to tighten the fasteners, and reduce it if possible (see TORQUE, SELECTION OF) Provide better training and/or supervision for the mechanics involved in the assembly Use more accurate bolting tools (see ACCURACY) Replace the fasteners with stronger ones of the same size. The following will increase their strength (see STRENGTH OF FASTENERS-Static, for details): -Use a new material having a higher strength -Use fine pitch instead of coarse pitch threads.

(NOTE: This increases body strength but decreases stripping strength)


o Change the design to allow larger fasteners, or more of them.

o Change the design to reduce loads on the fasteners so that you can reduce preloads.

FASTENERS, STRIPPED

(See THREAD STRIPPING)

FATIGUE FAILURE

General

It's a fatigue failure when:

o The failure is sudden with little or no necking-down of the part.

o The component has been subjected to cyclic tensile loads. n Usually the cyclic loads are well below the material tensile

strength.

Fatigue failures are most easily identified by the appearance of the fracture surface (beach marks, polishing, corrosion).

Typical appearance of fatigue failures are shown in the following Figure.


FIGURE 12 Cross-section of a bolt which has failed in fatigue. The beach marks are shown at (A). The rougher surface, where the boit failed abruptly, is shown at (B).

Whether or not a part will fail in fatigue depends upon many factors, including the properties of tlie material from which the fastener is made, the way it was processed, defects in the material, stress levels, details of the shape of the fastener (thread radii, for example). It takes an expert to determine why a part failed.

Analyzing and Reducing Fatigue Problems Fatigue is a complex subject involving many variables. If your problems are frequent and severe or you are especially concerned about the consequences of a fatigue failure, you should consult an expert. If you have to solve the problem yourself, however, here are some of the things which have helped others:

Protect The Evidence

You may not be able to solve the problem. If so, you'll be glad that you have saved samples to show to an expert. Place them in a bag to protect them from the environment. Here are some of the things a broken bolt can tell the expert:


o Where and when failure started-quench crack, inclusion, pit.

o Beach marks show a time history of the crack progresion. Corrosion products and polishing of surface give clues.

o Final overload. This can give an indication of the external load to which the fastener was subjected for example (final overload area was 10% of section; therefore, the external cycle load was small, approximately 10% of tensile strength).

Increase the Preload

Rapid fatigue failure can result if preloads are too low. If preload is low, the bolt will often see a larger portion of an external load. PRYING action is often increased too, so your first step in combating fatigue should be to recheck your preload and/or torque specifications (see PRELOAD, SELECTION OF).

Review assembly tools and procedures to make sure you're getting the specified preloads. And don't forget to consider RELAXATION effects when reviewing your preload specifications and assembly procedures. It's the residual preload, not the initial preload that matters when the fatigue (cyclic) loads are applied to the jOint (See PRELOAD, INITIAL/RESIDUAL).

If you believe you're getting the specified preloads and they're not doing the job, increase them. Preloads as high as 70-100% of yield have been recommended in some applications (see PRELOAD, SELECTION OF-Level 3 or 4).

Inspect the Surfaces

Fatigue cracks can start at minor defects, pits, cracks, or folds in the surfaces of the bolt. Inspect the surfaces of the bolt. Discard any with defects.

Avoid Decarburlzation

A decarburized surface is easily "cracked" and encourages fatigue. Make sure bolts have been properly heat treated.


Perpendicularity

Check the perpendicularity between nut face and the axis of the baIt. Tests show that a nut face-to-bolt-axis angularity of only 4 degrees can reduce fatigue life to almost nothing.

Possible solutions include spot-facing or milling the flange surface, use of tapered washers or a self-aligning nut (Fig. 13).

FIGURE 13 Cut-away view of a two-piece spherical nut showing how it conforms to a joint surface Which I .. not perpendicular to the bolt axis.

New Bolts

Periodically replace the bolts, before they fail-a simplistic, but effective way to avoid fatigue failure; since fatigue failure is time dependent.


Materials

Make sure you're using a fatigue resistant bolt material. Consult your fastener suppliers for advice. Some of the material properties often cited as desirable when fatigue is a problem include:

o Relatively high strength, but not too high. An ultimate tensile of 160 ksi is considered good for fatigue resistence. Check the stress corrosion cracking properties of the material (STRESS CORROSION CRACKING).

o At least 7% minimum elongation o Through hardenable o High notch strength

One bolt material which meets these criteria is 4340 heat treated to 160 ksi tensile strength (145 ksi yield).

Nut and Bolt Shapes

When you're talking to your fastener supplier about material, also ask about fatigue resistant shapes. Little differences can have a big effect on fatigue life. The radius of thread roots, for example, has a major impact on life (large, smooth radii giving better life than sharp roots). Rolling the tlueads (especially after heat treatment) instead of machining them can be important, too. The shape of the fillet between head and body, the way the threaded section is blended with an unthreaded body, the shape of the nut and bolt head-these and many other factors can make a difference in fatigue life. Good, fatigue resistent fasteners are available.

Nut shape and details are equally important. Flanged nuts with tapered threads on the inboard face, longer nuts, can all make a difference. Again, see your fastener supplier for suggestions.

Stress Concentrations

Eliminate unnecessary stress concentrations. For example, if thread run-out is too close to the nut bearing surface, the thread stresses at the first thread of engagement of the nut will be inceased. As shown below at least three full threads below the nut face and one or two sticking out beyond the nut should guarantee full engagement.

Bolting Procedures Reference Ma. 39

FIGURE 14 A minimum of two threads above the nut and three threads below will help improve fatigue life.

__ -L~=~\.,;:=:= 3 THREADS

Lock nuts can also have a beneficial effect on the distribution of thread stress. So can jam nuts or helicoil inserts and tension nuts.

FIGURE 15 Cut-away view of a tension nut.

40 Bolting .:edures Reference Manual

Fatigue Tests

Because of the number of variables involved, the fatigue life of fasteners cannot be predicted. Tests made on conventional polished test coupons help to evaluate a given material; but the actual life of a bolt made of that material is usually only 1/2 to 1/4 the life of the test coupon, because of stress concentrations created by threads, head-to-body fillet, defects caused by processing, and environment.

If you have a problem, it may be necessary for you to have a qualified laboratory conduct tests on your actual bolts under conditions that simulate your application as closely as possible.

Reduce the Applied, Cyclic Loads

It's usually impossible in a field situation to reduce the loads seen by the bolts. These loads are a function of operating temperatures and pressures, equipment design, and joint design. If all else fails, however, you may have to ask the equipment designer to do this for you. There are loads, called endurance limits, below which ferrous materials will have an infinite fatigue life. These loads depend upon the shape of the fastener, its condition, as well as on the material. Unfortunately the endurance limit of bolts is very low-perhaps only 30r 4% of the yield strength of the material-so it's unlikely the designer can give you this much relief. But the closer he comes to the limit, the longer the bolt will last.


1. Fatigue Design Handbook, published by the Society of Automotive Engineers, AE-4 Advances in Engineering Series No.4.

2. Bickford, J.H., Section 23-Bolted and Riveted Joints. Standard Handbook of Machine DeSign. Editors: J. Shigley and C. Mischke, McGraw Hill, New York. 1986.

FINISH, FLANGE

(See LEAKS)

------------------ ..... \--Bolting Procedures Reference M, . 41

FLANGE, CAST IRON

(See FLANGE, UNCOMMON MATERIALS)

FLANGE, OVAL RING

(See GASKETS, METAL a-RINGS)

FLANGE ROTATION

The outer edges of a raised face flange are pulled towards each other when the bolts are tightened. This distortion is called flange rotation.

FIGURE 16 Flange rotation. Note how the gasket Is compressed near the bolt and relieved towards the Inside of the vessel.

(

42 Boltine cedures Reference Manual

FLANGE, STAINLESS STEEL

(See FLANGE, UNCOMMON MATERIALS)

FLANGE, UNCOMMON MATERIALS

Recommendations for TORQUE, PRELOAD, GASKET STRESS, etc., are usually based on the assumption that the flanges involved are made from carbon steel. When they are not, you must take extra steps to avoid possible problems.

Strength of the Flange

Many materials are not as strong as carbon steel. Applying normal torques can Yield them. For example, here are some typical yield strengths for flange materials:

Carbon Steel (SA 105) Stainless Steel (SA 185)

100 po

36 ksi 30

500 po

29 ksi 19

1000 po

20 ksi 16

As a result, one utility reduces the maximum torque applied to stainless steel flanges as follows:

Operating Temperature

Below 200 po 200-500 po 500-700 FO .47 T

Torque

.87 T

.60 T

Where: T = Torque for carbon steel flange

Gray iron fl~nges are used for low temperature and low pressure service. Matenals such as ASTM A126 don't have a defined yield strength. The tensile strength of A126 is:

ASTM A126 Gr A ASTM A126 Gr B

20 ksi 30-31 ksi

Sedion VIII of the ASME Code requires a factor of safety of 10 be apphed to these tensile values to arrive at an allowable stress. This makes it almost impossible to show that raised faced gray iron flanges meet code stress requirements.

Bolting Procedures Reference·" \pal 43

One utility used 30 ksi bolt loads for full faced gray iron flanges. They broke a cast iron 125# flange at these bolt loads by bolting it up to a steel raised raced flange.

Thermal Stress

Uncommon flange materials may have coefficients of expansion that differ substantially from the coefficients of the bolt materials. The coefficients may vary with temperature (see COEFFICIENT OF EXPANSION for additional data).

Material

Carbon Steel Cast Iron Stainless Steel

Coefficient of ° Expansion (in/in/F) @ 600 deg. F

8.35 X 10' 6.20 X 10' 10.38 X 10'

A temperature rise will cause an increase in bolt preload and in the clamping force on the joint if the bolts are low alloy steel and the flange members are stainless steel due to the difference in coefficient of thennal expansion. See THERMAL STRESS for further information, and for procedures for calculating such stresses.

GALLING (See Also LUBRICANTS) Galling is the cold welding, or partial welding, of one heavily loaded

metal surface to another. It is encountered when those surfaces are brought together so intimately that molecular bonds adually form between mating parts-between a bolt and its nut, for example.

We see this intimate contad when surfaces are highly loaded, when lubricants have migrated or dried out, when we have damaged threads (to create stress concentrations), tight thread fits, and or combinations of such things. Corrosion and high temperatures can make the prob

lems worse.

44 Boltirlo Procedures Reference Manual

To Avoid Galling

There is no known way to eliminate all 11" dally on larger fasteners Experie h ga mg of bolts and nuts, espe-ing things can help Y~u'll h nce s ows"however, that the follow-

I. . ave to try these thi .

app 1cations to learn which are eff t' f ngs m your own ec lve or you.

D Use coarse threads instead of f o T . me. ry vanous anti-seize compounds Bef h .

cants, ensure that the lubricant is 'com~:~i~I:~~~gt~Ubri-component (see LUBRICANT) H e we've heard in the field: . ere are some comments

-Molydisulphide anti-seize works well if bolt I d than 50% of yield. oa s are less

-Fel-Pr~ C670 is good, even on rough surfaces when operatmg temperatures are less than 750 d ' F L' 'd (d' h egrees - T~U1 1S) detergent is effective on aluminum .

- . e best anll-galling lubes are silver-based . -~;Ik of magnesia is an effective anti-gallin~ lubricant - s:e~lo~~I~r~~se tgives dgool d galling protection for stai~less

a appe a UIDlnum hole.

D ~:I~i~:.t Fa~~~~~;::terials which are somewhat resistant to

-400 series stainless nuts work well on 316 . b I ARMCO N't . ser1es a ts - N' . 1 romc 60 stainless seel bolts work well wi;h

'trond,c 50 nuts (but not, surprisingly, the other way

aroun ). -It helps to use stainless steel nuts on B7 B16 d th

low alloy steel bolts. ' an a er -Cold drawn 316 works well on cold drawn 316.

Removal of Galled Studs and Bolts

Just as there are no magic ways to prevent all' ways to remove seized studs or bolt H g mg, there are no perfect have found helpful h s. ere are some tncks that others

, Qwever:

D ~St iodine as a p~netrating oil. Be sure to clean it off the

bo ts, nuts and Jomt members after removing the bolts ecause 1t's a mild add d '11 ' an W1 continue to penetrate the

parts. e] ~ther mild acids can also be used as penetrating oils with

e same ~auhons as above. '

Bolting Procedures Reference '~\Ual 45

o Sometimes heating andlor cooling a bolt will break it loose. Drill a heater hole in the bolt if necessary.

o Weld a nut to the end of a stud; use the nut as a "head" to get a good grip on the bolt. Add penetrating oil, and untorque the stud. (It sometimes helps to heat the flange a little, too, if allowed.)If they break, drill them out.

D Drill a hole through the flange or drill the stud to reach a blind hole at the far end. Tap the outer end of the hole; add a pressure coupling; and pump penetrating oil into the blind hole under pressure (2000 psi has been used). Maintain the pressure awhile, then untorque them.

o If all else fails, drill them out using a magnetic base drill or an EDM machine. The latter leaves a heat-treated skin on the hole and this can be a problem when you try to re-tap the hole. After drilling, try peeling out the threads from the bolt; this will save the tapped hole.

o As a last resort, the hole may have to be drilled oversized and tapped to accept an insert. This is a configuration change, and design of the component will have to be evaluated for this change.

GASKET, CREEP OF Virtually all gasket materials are partially plastic in their behavior;

they reaIly must be in order to conform intimately with the minute inegularities in the joint surfaces and therefore to prevent leakage. The fact that they're not fully elastic, however, means that the whole gasket will creep and flow to some extent when placed under heavy loads.

The gasket gets thinner and wider as a result of this creep, allowing the bolts to relax a little. Stresses on the gasket are relaxed and a leak path may open, even if the joint was tight when initially assembled.

It is widely believed that this creep accounts for the often gross loss of preload found in the bolts of a previously tightened joint-the loss which occurs between passes, for example. Recent tests by the Pressure Vessel Research Committee's Task Group on Gasket Testing, however, show that the creep of most gaskets-spiralwound, compressed asbestos, asbestos substitute, double-jacketed, etc.-is smaller than commonly supposed. Preload losses caused by creep rarely exceeded 5% of initial preload, at least during room temperature tests, and the bulk

46 BoIting P,.. .ures Reference Manual

of the loss occurred within minutes of initial loading. It is now believed that ELASTIC INTERACTIONS are the principal cause of loss of preload in gasketed joints tightened in multiple passes. See PRELOAD, LOSS OF for further details.

Two exceptions to all of this are elastomeric or plastic gaskets, which do exhibit a Significant amount of creep and which will continue to creep for several days after tightening. (See GASKETS ELASTOMERIC OR PLASTIC)

GASKET LEAKS

(See LEAKS)

GASKET STRESS

(See LEAKS; GASKETS, IN GENERAL) Whether a gasketed jOint will leak or not depends upon the clamping

pressure exerted by the flange members on the gasket. These surface pressures are called gasket stress.

Experiments and experience show that two different gasket stresses are important:

o The initial seating stress created when the bolts are first tightened and

11 The residual stress on the gasket after the system has been pressurized (partially relieving the gasket in most cases).

The initial seating stress on the gasket is related to the "y" gasket factor used for design purposes in the ASME Boiler and Pressure Vessel Code. Table XI-3221.1-1 in Section III, Division 1 of the Code lists suggested (not mandatory) y factors for a variety of common gasket materials. For a spiral-wound, asbestos-filled gasket, the Code recommends an initial seating stress of 10 ksL

The residual stress on the gasket is related to the" m" gasket factor also tabulated in Table Xl-3221.1-1. This m is a "maintenance" or "multiplying" factor. It defines recommended, residual gasket stress in

Bolting Procedures Reference Ma/~-) 47 I

terms of some multiple of the pressure contained by the system. For a spiral-wound, asbestos-filled gasket, for example, the table gives a recommended m of 3. Given a contained pressure of 1,000 psi, therefore, the residual stress on the gasket should not fall below 3 ksL

The Code recommendations are complicated by the fact that they are not based on the full contact surface area of the gasket. The original authors assumed that FLANGE ROTATION would unload the inner portion of the gasket, if the flange had a raised face. They recommended, therefore, that only about half of the gasket surface be assumed to be in contact with the flange members when the seating or residual stresses were evaluated. So a recommended yof 10 ksi, computed on the half area of the gasket, equates to an actual average seating stress of only 5 ksi for the full area.

Recent studies, sponsored by the Subcommittee on Bolted Flanged Connections of the Pressure Vessel Research Committee, suggest that it is usually more accurate to assume full contact between the gasket and the flange. The Subcommittee is in the process of preparing revised recommendations for gasket factors and the new ones, it is hoped, will be based upon full gasket area.

Table C gives some suggested seating stresses for com~on gasket materials, based on full gasket area. These recommendatlOns are the result of the PVRC studies mentioned above. As a result, they differ from current Code suggestions. Will you violate the Code if you use them?

The answer is, "No, you will not." Remember, the Code values are suggestions only and are intended for use by designers; they are not mandatory, and are not intended to define assembly or in-service gasket stresses.

In fact, the Code does not give very specific information about the assembly of bolted flanges. Only Appendix 11 of Section III (or the identical Appendix S of Section VIII, Division 1) deal with assembly stress. Both, in effect, say that you should tlghten the bolts enough to prevent leaks. And the reader is told that this will mean bolt stresses in excess of the "allowable" stresses given in Code Table 1-7.3 and 1-8.3.

The Code is not contradicting itself when it says this. The allowables are intended to be used for design purposes only. They force the designer to overdesign the joint; to be conservative and therefore safe. The bolts selected by this design process, however, now can:-and should-be tightened beyond the stress limits allowed the deSIgner. The Code lists possible tightening values (by suggesting y factors) to help the designer size the bolts and flange members. The PVRC studIes, however, suggest values which are better for assembly purposes.

48 Bolting Pro res Reference Manual

To be complete, we should mention that the PVRC work has shown that there's no such thing as a I'perfect" seating stress or maintenance stress for a given gasket. The "best" factors for a given application depend upon such things as the type of fluid contained by the system, the contained pressure, the flange finish, etc. The maintenance factor is also a function of the seating stress and vice versa. Increase seating stress and you reduce the maintenance factor required to contain a given fluid at a given pressure.

To further complicate things, even if you can determine which seating stress would be best for a given gasket, contained fluid, etc., you may not be able to use that stress when you tighten the joint. In some cases (for example, in small diameter, low pressure flanges) you'll find that the bolts would have to be stressed near or even past yield to develop the desired gasket stress. In large flanges you may find that the theoretically ideal bolt loads will cause excessive FLANGE ROTATION; and that this can actually increase the tendency to leak. In other words, a lower assembly stress, which pure gasket tests suggest would increase the leak rate, will actually decrease it.

THERMAL EFFECTS, which aren't welI understood at present, can also force you to use lower assembly stresses than those suggested by pure gasket tests. So can a concern for STRESS CORROSION CRACKING. In some cases you'll also be limited by the size of the wrench you can use On a given joint. Or by the stresses the bolts would create in the flange members. There are limits to how much stress a gasket can support, too (see the manufacturer). You should probably have a designer check your proposed assembly loads if you face severe thermal cycles and/or are planning to tighten the bolts past 40% of yield.

As a result of all this, it's not possible to give you a table of gasket seating stresses, or maintenance factors, which will be best for every situation. Trial and error will often be required if you have to deal with FLANGE ROTATION or THERMAL EFFECTS or the like. It IS possible to say, however, that seating stresses should, in general, be higher than the y factors used for design purposes. The following Table takes this into account. Note that the stresses given here are based on the full area of the gasket; not on the Code's reduced area. Note, too, as mentioned above, that you will often have to use stresses even higher than those in the Table, to seal a problem joint.

Bolting Procedures Reference r 49

TABLE C Minimum Seating Stresses for Gaskets (For Assembly Purposes, Not for Flange Design) Based on Full Gasket Area

Gasket Material

Self-energizing types (O-rings, metallic, elastomer, other gasketed types considered as self-sealing)

Elastomers without fabric or high percent of asbestos fiber: Below 75 Shore Durometer 75 or higher Shore Durometer

Asbestos with suitable binder for operating conditions: 1/8 in. thick 1/16 in. thick 1/32 in. thick

Elastomers with cotton fabric insertion

Elastomers with asbestos fabric insertion (with or without wire reinforcement):

3 ply 2 ply 1 ply

Vegetable fiber

Spiral-wound metal, Asbestos Filled: Carbon Stainless or Monel Graphite Filled: Stainless Steel Chlorite-Graphite Filled: Stainless Steel

Corrugated metal, asbestos inserted, or corrugated metal, jacketed asbestos filled:

Soft aluminum Soft copper or brass Iron or soft steel Monel or 4-6% chrome Stainless steels

Corrugated metal: Soft aluminum Soft copper or brass Iron or soft steel Monel or 4-6% chrome Stainless steels

Flat metal, jacketed asbestos filled: Soft aluminum Soft copper or brass Iron or soft steel Monel 4-6% chrome Stainless steels

Minimum Seating Stress psi

o

o 150

1200 2775 4875

300

1650 2175 2775

825

7500 7500 7500 7500

2175 2775 3375 4125 4875

2775 3375 4125 4875 5700

4125 4875 5700 6000 6750 6750

50 Bolting 1 ~vcedures Reference Manual

TABLE C Continued

Grooved metal: Soft aluminum Soft copper or brass Iron or soft steel Monel or 4-6% chrome Stainless steels

Solid flat metal: Soft aluminum Soft copper or brass Iron or soft steel Monel or 4-6% chrome Stainless steels

Ring joint: Iron or soft steel Monel or 4-6% chrome Stainless steels

GASKETS-ELASTOMERIC OR PLASTIC

4125 4875 5700 6750 7575

7500 7500 9000

10900 13000

13500 16350 19500

_ G.askets made ~f elastomeric or plastic materials can be especially ~iffIcult,;o deal with because they are dimensionally less stable than . harder gaskets such as compressed asbestos, spiral-wound, doubleJacketed, solid metal, etc. One company reports that they find it necessary to compensate for the creep of such gaskets by retightening them (~t rated torque) twenty-four hours after initial assembly and sometrmes agam after forty-eight and seventy-two hours.

GASKETS-IN GENERAL

(See GASKET STRESS; LEAKS; GASKETS, METAL O-RING; GASKETS, METAL)

---------------------~ ~.\--~

Bolting Procedures Reference. ;lal 51

Basic Considerations

Gaskets are used to fill irregularities in joint surfaces -to plug the gaps between joint members-and so to prevent leaks. To do this job, gaskets must be:

o Soft enough to conform completely with both large and small imperfections in joint surfaces_

o Elastic enough to follow the joint members as they expand and contract under pressure, thermal and/or other loads.

o Able to resist chemical attack by the contained fluids, at operating temperatures.

o Able to accept the high clamping forces required to mate the gaskets intimately to flange surfaces (without, for example, excessive cold flow or creep).

Selection of a particular gasket, for a particular job, is a task for the designer or for the company which manufactures gaskets. Companies such as Union Carbide, Lamons and Flexitallic have all published technical manuals which give you a great deal of information about gasket choices and the proper use of gaskets. You should request and read these manuals if you have to pick a gasket.

Our purpose in this reference manual is simpler-to help you understand and troubleshoot gaskets already selected by the designer. As in all other aspects of bolting, the designer cannot guarantee success; the people who assemble and maintain the joint must do their part correctly or the joint will still fail .

Use and Installation of Gaskets

Here are the factors which the field people must control:

o The condition of the gasket. Don't re-use them. Handle them carefully during assembly. Bent, nicked, gouged or hammered gaskets won't seal.

o The condition of joint members. Even a perfect gasket will be unable to seal badly damaged or warped flange surfaces. Radial gouges, scratches, tool marks, etc., are especially bad. The finish of flange members is one aspect of their condition which deserves attention. The flbest" finish for a particular type of gasket is a hotly debated subject. Many companies have rigorous specifications on this. Recent studies by the Task Group on Gasket Testing of the Pressure Vessel Research Committee, however, suggest that

52 Bolti! 'ocedures Reference Manual

finish is not as critical a factor as many think-except perhaps for solid metal gaskets. The leakage behavior of spiral-wound, compressed asbestos, asbestos substitute, metal-jacketed asbestos, and oval ring gaskets was found to be relatively unaffected by variations in gasket finish ranging from 32 inches to 1000 micro inches. Nevertheless, the debate still rages and a lot of people with a lot of practical experience dispute the PYRC conclusions. See the gasket manufacturer for recommendations.

Installation

Joint integrity depends in good part on satisfactory gasket installation. Here is some practical guidance:

o The placement of the gasket. It must be centered on the joint members, clear of the bolts, etc.

o The initial preloading of the gasket. The joint must be drawn together carefully. One plant claims that unless flanges are parallel within 0.012" before tightening the bolts, the jOint will leak. Most people don't hold parallelism to this tight a tolerance, but it certainly is important.

o Initial loading. The gasket must be loaded uniformiy (within reason); never tighten one bolt fully and then move on to the rest. Tighten them partially in several passes in a crossbolting pattern (see ASSEMBLY PROCEDURES) to compress the gasket uniformly. .

o The most important-and most difficult-gasket job facing the field worker is to produce the correct preload in the joint. The initial seating stress on that gasket must be low enough to avoid damaging the gasket, but high enough to prevent a significant leak. See GASKET STRESS for further information.


1. Gasket Handbook. Lamons Gasket Co., Houston, Texas 2. Flexitallic Spiral-Wound Gaskets, General Catalog.

Flexitallic Gasket Co., Inc., Camden, New jersey. 3. ASME Boiler and Pressure Vessel Code, Section III,

ASME, New York, 1984.

----------------- '-, ---Bolting Procedures Referenc£ \ual 53

4. Payne, J., A. Bazerguiand G. Leon. New Gasket Factors-A Proposed Procedure. ASME Pressure Vessel and Piping Conference. New Orleans, june 1985.

5. Robinson, j. N., M. l. Lundin, and l. Spiewak, Development of Ring-Joint Flanges for Use in the HRE-2, ORNL-3165, Oak Ridge National Laboratory, Oak Ridge, Tennessee, December 1961.

GASKETS, METAL O·RING

Most of the discussion of gaskets in this manual is based on the assumption that you're dealing with a raised face flange and the gasket is a type appropriate for that flange (spiral-wound or sheet asbestos or double-jacketed or the like). You're told, for example (under LEAKS) that if the joint leaks you should try increasing the initial preload on the joint. You're given suggested seating and residual stresses for various gaskets (under GASKET STRESS) and told to use these values to compute an appropriate preload, etc.

Joints with metal O-ring gaskets don't obey the rules for raised face "normal" joints. Instead, for metal O-rings:

o The initial and residual gasket stresses required are very low.

o The bolt load required to compress the gasket and achieve metal-to-metal contact of the joint depends upon the diameter of the gasket, the material and its wall thickness. Seating force or stress values can be obtained from the gasket manufacturer.

o The bolt load required to assemble such joints is equal to the sum of the small force required to seat the gasket, plus the larger force needed to prevent the joint from separating under the internal pressure load:

FB = (FG + AG P) I n Where FB = required assembly bolt load

FG = force reqUired to seat the gasket A G = area of gasket

n = number of bolts P = internal pressure

54 Bolting,.' edures Reference Manual

The performance of the metal a-ring gasket is relatively insensitive to flange finish. One manufacturer, for example, recommends a finish of 32 micro-inches, but says that the gaskets will seal at finishes up to 250 micro-inches, "especially if the gaskets are heavily plated".

The same manufacturer also offers some advice for troubleshooting such joints:

If the joint seals for any length of time, and then starts to leak or if it seals at first, but leaks before maximum contained pressure has been reached, you probably have a bolting problem (insufficient preload).

If it never seals, it's a bad gasket or bad surface finish, damaged surface, or warped flange.

Note that increasing the bolt preload in the latter case will only increase the interface contact pressure between joint members. It will not increase gasket stress or cure the leak.

FIGURE 17 A flanged joint with a metal O-ring gasket. Contact is basically metal-to-metal.

Bolting Procedures Reference'~ "uaI 55

GASKET, METAL (SOLID OR CORRUGATED)

(See GASKETS IN GENERAL) Solid or corrugated metal gaskets don't behave like more common

ones (such as spiral-wound, compressed asbestos, doublejacketed, etc.).

For example, the leak behavior of most common gaskets is improved if the preload applied at assembly is increased a little (see LEAKS or GASKET STRESS). That's not true of metal gaskets; extra preload often won't cure a leak.

The behavior of most common gaskets is becoming predictable, thanks to the work being done by the Subcommittee on Bolted Flanged Connections of the Pressure Vessel Research Committee. The experimenters, at least, now know how to evaluate a sample spiral-wound gasket, for example, and how to use their tests on the sample to predict leak rates under a wide variety of conditions. They report, however, that the behavior of metal gaskets is "very erratic and unpredictable, especially when they're used on freshly machined flanges':. The b~havior becomes more predictable when a new gasket IS used In a prevIously used flange. Even he:e, however, they don:t leak at a slowly increasing rate as pressure mcreases the way.a spIral-wound gasket will, for example. Instead, the metal gasket WIll seal for awhile-and then suddenly start to leak.

The leak rate of a spiral-wound gasket is affected by flange surface finish; but only a little. The behavior of a metal gasket is strongly affected by surface finish (see the conunent above on new versus reused flanges, for example). Consult the manufacturer for recommendations.

GRIP LENGTH The distance between the inner face of the nut and the underside

of the head of the bolt (or inner face of the other nut) including washers, if any. The combined thickness of all the parts which are clamped together by the bolt. If a tapped hole is used, instead of a nut, the member containing the hole is considered to be a /I nut"; a clamping part rather than a clamped part.

58 ~ . .....--

Ig Procedures Reference Manual ,j~ Bolting Procedures Refe: Manual 59

" :a

TABLE D ~~ Continued

~!~ Measuring the true hardness of a bolt in the field can be a tricky thing

ASTM Type to do. Surface measurements can be misleading, because of such factors

1/2" to 33-38 HAC

;J~ as localized work hardening, surface decarburization, rust and surface 490 1,2 1 1/2/1 592-628 150 Mini

ASTM B21 Under 2" 170 Max damage. As a result, the proper way to measure the hardness of a bolt

540 CI5 241-285 HB 520-562 120 wd.:J:ij is to cut a piece off one end, and to check the hardness at several radii

2-6/1 248·302 HB 528·578 on the exposed crosssection. The section tested must be at least one

6-8" 255-211 HB 115

~ diameter from the end of the fastener.

B21 Under 31/ 534·490 115 CI4

269·331 HB 548·604 135 Since it is often not possible to cut off the ends of bolts in field situa-3-6" 277-352 HB 556·620 I tions, the Joint Task Group on Bolting of the AIF and MPC sought and

B21 Under 3/1 293·352 HB 135

Ir·~ found a surface measurement procedure that can be used to give you CI 3 570-620 145

3_6" 302·375 HB ~ a rough indication, at least. The method has been defined in a document

B21 Under 4/1 • 578·640 145 called STANDARD TEST METHOD FOR EQUOTIP R HARDNESS CI2

311-401 HB 586·660 155 TESTING OF METALLIC MA TERlALS, previously distributed by EPR! B21 Under 4 11 ~ Cll

321-429 HB 594·680 165 to utilities and submitted to the ASTM with the recommendation that B24 Under 6/1 248-311 HB 528-586 ~

it be adopted as a standard procedure. Hardness is defined by anL CI 5 120 (Leeb) number. The document referred to includes Tables of conversion

6/1-8" 255·321 H8 factors between Leeb, Rockwell and other hardness standards. Leeb 8"·9 534·594 115

~ 1/2/1 262·321 HB 542-594 115 hardnesses are also included in the Table above.

824 Under 3/1 269·341 H8 548-612 ~ Further information about the Equotip Hardness procedure (and test

CI 4 135 equipment) can be obtained from: 3/1·6" 277-352 HB

6/1·8" 556·622 135

.,..A~ 8"·9 285-363 HB 564·630 135 Hentschel Instruments, Inc.

1/2/1 293·363 HB 570·630 135 2505 South Industrial Highway

B24 Under 311 293-363 HB 570·630 ~ Ann Arbor, Michigan 48104

CI3 145 Tel: 313/973-2505 3//-8" 302·388 HB 578·650

~~ 8"-9 311·388 HB 145 J.M. Devine Company 1/21/ 586·650 145

B24 Under 7/1

311-401 HB ~ 716 Washington Street

CI 2 586·660 155 P.O. Box 307 7"·9 321·415 HB 594-670 Dedham, Massachusetts 02026 1/2" 155

~ B24 Under 6/1 321·415 HB Tel: 617/329-7778

Cll 594·670 165

~ 6"·8" 331-429 HB 604·680 165 SAE GA 2 1/4" 10 80·100 HAB J429 3/4/1 414·516 74

~ Over 3/4/1 70·100 HAB 380-516 to 1 1/2" 60

~ GA 5 1/4" to 25·34 HAC 534·594 1" 120 HEATERS Over 1/1 19·30 HAC 502·572 ~ to 1 1/2/1 105 One common way to tighten large bolts is to insert a heating rod

GR 8 1/4" to 33-39 HAC 1 1/211 592·634 150 in a hole drilled down through the center of the bolt; stretch the bolt

~ by heating it; run the nut down against the joint; and remove the heater. HRS = Rockwell "8"

~ HRC '" Rockwell "e" As the bolt cools it shrinks, developing tension or PRELOAD. HB = Brinell

;!~ F·'i.~

56 Bolth 10cedures Reference Manual

FIGURE 18 An Illustration of grip length (LG). Note that the grip length Includes all clamped parts, Including a washer) If any.

~--lG >

HARDNESS OF FASTENERS

There's a rough correlation between the hardness of a fastener and its strength. It's sometimes helpful, therefore, to measure hardness to help identify unmarked or s.uspect bolts in the field. The Table below gives specified or expected hardness ranges for common bolting materials. Procedures for measuring this hardness, in the field, are given after the Table.

TABLE D Hardness-Ultimate Strength

Spec.

ASTM 193

Grade or Class

B7 B16

B7M B8T B8 B8C

ASTM L7 320 L7A

L7B L7C L43 B8 B8C B8M B8F B8T

ASTM Type 325 1,2,3

ASTM GR A 307 GR B

ASTM BC 354

ASTM 449

BD

Size

2 1/211 and under 21f2"·4" 4-7"

4"& undo

1/2"-111 1 1/8/1-1 1/2/1

1/4" to 21/2/1 Over 2 1/2/1 1f41f to 2 1/211 Over 2 1/2"

1/4'1 to 1" Over 111 to 1 1/2 1[

Over 1 1/2" to 3/1

Bolting Procedures Reference M '~~?l 57

Hardness (Rockwell Leeb or Number Ultimate

Brinell) (L) (ksi)

269·275 HB 548·556 125

244·252 HB 524·532 115

232·236 HB 512-516 100 99 HRB 508·512 100

96 HRB 488·492 75

96 HRB 488·492 75

96 HRB 488-492 75

269·277 HB 548-556 125

269·277 HB 548·556 125

269·277 HB 548·556 125

269·277 HB 548·556 125

269·277 HB 548·556 125

96 HRB 488·492 75

96 HRB 488·492 75

96 HRB 488·492 75

96 HRB 488·492 75

96 HRB 488·492 75

24·35 HRC 534·606 120

19·31 HRC 510·578 105

69·100 HRB 378·516 60 Min

69·95 HRB 378·486 60 Mint 100 Max

26·36 HRC 546·614 125

22·33 HRC 522·592 115

33·39 HRC 592-636 150

31·39 HRC 578·636 140

255·321 HB 534·594 120

(25·34 HRC) 223·285 HB 502·564 105

(19·30 HRC) 458·516 90 183·235 HB

60 Bolting F dures Reference Manual

The method is relatively slow but it's inexpensive (heaters are cheaper than high torque tools, for example). By itself, heating is not an accurate way to develop or control a specified preload. If DATUM RODS and depth mics or ULTRASONICS are used to measure the residual STRETCH of FASTENERS, however, the method is very accurate. (See STRETCH OF FASTENERS AND JOINTS).

There is some danger that heaters can decarburize the surfaces of bolts, weakening them and leaving them more susceptible to FATIGUE and STRESS CORROSION CRACKING. They can be useful tools when GALLING is a problem, however, since you avoid high contact forces between moving parts.

Disassembling bolts which have been heated can be a problem. Presumably you'll heat them to relieve the preload. Now you often must use a torque tool of some sort to disengage threads which have frozen together because of CORROSION, EMBEDMENT, etc. You still won't need as large a wrench as you would for loosening the bolts. "cold", however.

Procedures for heating bolts should include the following:

o Heat several at once at cross points around the joint to minimize ELASTIC INTERACTIONS.

o Go for final stretch (preload) in a single pass. o Use depth mics or dial gages with DATUM RODS or

ULTRASONICS to measure the residual preloads after the bolts have cooled.

o Reheat and retighten those which aren't right (skilled operators claim to be able to get about 60% of them right on a first pass, 30% on a second pass, and the rest on pass #3.

IDENTIFICATION OF MATERIALS

(See MATERIALS, IDENTIFICATION OF; HARDNESS OF FASTENERS)

Bolting Procedures Reference Me 61

INSPECTION OF BOLTED JOINTS

Many techniques have been developed for inspecting welded joints. Unfortunately very few are available for the equally important (and complex) bolted joint and most of these give only an indirect evaluation. Here, for example, are some of the choices:

Residual Preload Tests

Whether or not the joint is going to perform its intended function depends to a large extent upon the preload in the bolts and the resulting clamping force on the joint interface. There's no way at present to walk up to a previously tightened joint and tell with any great a~curacy h?w much preload is in the joints (at least without essenhally disassemblmg the joint in the process). But the following tests can give you a crude idea of the preload:

Torque Measurement Test Measure the torque required to restart the nut in a clockwise direction. Then use the familiar equation T ~ K D Fp / 12 to compute the apparent preload (See TORQUE, SELECTION OF). For nut factor K, use a value taken from the NUT FACTOR Table if the joint has been freshly tightened. It's probably more accurate to use a K that is 20% higher than the nommal tabulated value if the joint was assembled several months earlier and/or has been subjected to high temperatures (150 degrees F or higher). . .

The resulting estimate for preload, Fp, is probably accurate wlthm + / -50% in most cases. If the fasteners are rusty or galled or otherwise in poor condition, the accuracy will be less.

Other Tests Use DATUM RODS or ULTRASONICS to measure the reduction in length of one or a few bolts in the joint ~s you loosen the,;". Loosen them fully, one a a time, and carefully rehghten each-to. Its original preload, whether or not this was "correct" -before loosemng and measuring the next.

If a small sample suggests that the bolts in the joint are preloaded correctly, you don't need to do the rest. If there appea~ to be proble~s, you may want to test more and/or retighten all of them m a cross boltmg pattern to the desired preload.

Be sure to include both odd and even numbered bolts in your sample, since ELASTIC INTERACTIONS will typically leave a sawtoothed pattern of residual preload in the joint.

62 Bolting Pr;/--- U.res Reference Manual

Bolt Quality

If, after tightening the joint or after it has been in use for a while, you are concerned about the quality of the bolts, you can:

o Use a visual inspection for cracks, rust, damaged threads, etc.

o Test the hardness of bolts to estimate their tensile strength. You'd presumably do this, for example, if you became suspicious that the bolts were not made of the correct material, or had not been tested upon receipt.

o Use ULTRASONICS to test for possible cracks or wastage. Techniques for doing this have been developed and are described by S.N. Liu.

Reference The follOwing documents and texts have helped us prepare this section-and can give you additional information.

1. S.N. Liu. Status of Bolting Inspection. Paper D6/6 in Transactions of the 8th International Conference on Structural Mechanics in Reactor Technology, Brussels, August 19-23, 1985.

LEAK RATE

(See LEAKS) Studies by the Subcommittee on Bolted Flanged Connections of the

Pressure Vessel Research Committee show that there's no such thing as a totally "leak free" joint. There are, however, plenty of jointsmost of them-which leak so little that they're leak free for all practical purposes.

The results of the PYRC tests will be widely discussed and will be incorporated into the ASME Code in some form or other. Since these tests define gasketed joint behavior in terms of specific (non-zero) leak rates, a few words about rate are pertinent for this reference manual.

What's an acceptable leak rate? That depends upon the type of service, the location of the joint, etc. In many places, for example, a rate of one drop of water an hour would be acceptable. Such a rate would be unacceptable, however, if the joint which leaked was located over rOUT desk.

Bolting Procedures Reference M/ ~m""")I 63

The PYRC Subcommittee proposes three classifications for leak rates that are acceptable:

o Economy-rates up to 15 mg/s.mm are acceptable o Standard-rates to 1/500 mg/s.mm are acceptable o Tight-rates to 1/5000 mg/s.mm are acceptable

These rates are expressed in terms of mass leakage, instead of volumetric, because mass rates will define the behavior of a given jOint for all types of fluid (both liqUidS and gases) while volumetric leak rates are fluid specific.

Note that the rates are also expressed in "per mm" terms-this is "per mm of the flange's nominal diameter". The total leakage from a 9 inch flange is expected to be 50% greater than that from a 6 inch flange for the same conditions of contained pressure, initial and residual gasket stress, same contained fluid, etc.

The conversion tables below will help you convert these PVRC leak rates into terms that are easier to visualize. Note the significant impact that type of contained fluid has on the relationship between mass and volumetric leak rates.

TABLE E Leak Rate Conversion Tables

Class

Economy Standard Tight

Tightness Classificaflon



Mass Leak Rates mg/sec.mm Ibs/hr.inch

0.2 0.002 0.00002

Volumetric Leak Rate

0.04 0.0004 0.000004

cclsec.mm Pintslhr.in.

O.2x10-3 O.2x10-5 O.2x10-7

0.16 0.16x10-2 O.16x10-4

0.04 O.04x10- 2 O.04x10- 4

31 0.31 0.0031

Ruid

Water Water Water

Nitrogen Nitrogen Nitrogen

64 Bolting Prof ,-- 'res Reference Manual

The table below might also help you visualize various rates.

TABLE F Leak Rates and Bubble Equivalents

Leak Rate, Std. eelsee Volume Equivalent Bubble Equivalent

10- 1 icc/10 sec Steady stream 10- 2 icc/i00 sec i0/sec 10-3 3cc/hr 1/sec 10- 4 icc/3 hr 1/10 sec 10- 5 1cc/24 hr No bubble equivalent 10 - 6 1 cc/2 wk No bubble equivalent 10- 7 3cc/yr No bubble equivalent 10- 8 1cc/3 yr No bubble equivalent 10- 9 icc/3~ yr No bubble equivalent

Adapted from NUREG/CR·1312, UCRL·51738, Table 1. page 2

LEAKS

(See LEAK RATE) Leaks are one of the most common, and most troublesome, types

of joint failure. Various aspects of leakage are discussed in a number of sections of this report. This discussion, proceeding from simple to complex, will serve as a guide to those other sections.

Special Situations

When faced with a leak, you should first determine whether or not the joint involved is a raised face, carbon steel flange with a spiralwound or other common gasket. If so, proceed to the next paragraph. If not, then you should first read any of the following sections which are pertinent before proceeding to the next paragraph:

[) FLANGE-UNCOMMON MATERIALS [J GASKETS-ELASTOMERIC OR PLASTIC o GASKETS-METAL O-RING o GASKETS-SOLID METAL

/', Bolting Procedures Reference M jl 65

Review Assembly Procedures

The next step, for any type of joint or gasket, is to reexamine your boltup procedures. Are you taking the precautions normally required for gasketed joints? See ASSEMBLY PROCEDURES, GENERAL-All Joints, ASSEMBLY PROCEDURES-Torque, and GASKETS-in General

One important question: are you really getting the initial preloads you want in those bolts? A great many factors are involved in the torque-preload relationship, and many of them tend to give you less preload than anticipated. See ACCURACY for further details.

Another early question: does the finish on your flange meet the recommendations of the gasket manufacturer?

Review Preload Requirements

If you're satisfied that your procedures are up to par but you're still having problems, you should review your preload specifications. If this is a relatively unimportant joint and the leak is more a nuisance than a danger or expense, then you should review PRELOAD, SELECTION OF through Level #3.

If the joint is important andlor temperature effects may be a problem, you should review PRELOAD-SELECTION OF through Level #4. This level will lead you to a computation of GASKET STRESS and a consideration of THERMAL EFFECTS.

Tool ACCURACY is also considered in Level #4. You might want to add hard washers (see WASHERS-PLAIN) or use a LUBRICANT to reduce the scatter in achieved preload.

Compensate for Relaxation Effects

The next thing to consider, if you need more, is the fact that a number of factors will cause a time-dependent loss of preload in a previously tightened joint. Such loss cannot in general be monitored with torquing, tensioning or other common assembly tools. You might want to consider ASSEMBLY PROCEDURES-Stretch andlor the use of ULTRASONICS to monitor residual preload. See ELASTIC INTERACTIONS, GASKET CREEP and PRELOAD, LOSS OF for further details on the mechanisms which reduce the preload over a period of time.

66 Bolting Pro/ res Reference Manual

Sealants

You might want to consider SEALANTS as a next step in dealing with a chronic leaker. This is not an "elegant" solution, but it's effective and can get you by until you find a better answer.

Design Studies

The next level of attack is for the experts only. The design of the joint should be examined in detail. Things like FLANGE ROTATION should be estimated. The type of gasket, bolt loads, thermal effects, and the choice of materials should be reviewed. lf the design is faulty, you may have to take extraordinary steps to correct the problem-beef up or even replace the flange, for example. Or use something like ULTRASONICS to control the tightening process and compensate for ELASTIC INTERACTIONS.

References The following documents and texts have helped us prepare this section-and can give you ad.ditional information.

1. Gasket Handbook. Lamons Gasket Co., Houston, Texas 2. Flexitallic Spiral-Wound Gaskets, General Catalog.

Flexitallic Gasket Co., Inc., Camden, New Jersey. 3. ASME Boiler and Pressure Vessel Code, Section III,

ASME, New York, 1984. 4. Payne, J., A. Bazergui and G. Leon. New Gasket

Factors-A Proposed Procedure. ASME Pressure Vessel and Piping Conference. New Orleans, June 1985.

5. Robinson, J. N., M. 1. Lundin, and 1. Spiewak, Development of Ring-Joint Flanges for Use in the HRE-2, ORNL-3165, Oak Ridge National Laboratory, Oak Ridge, Tennessee, December 1961.

Bolting Procedures Reference M' 67

LENGTH, ACOUSTIC

The stretch of, or preload in, fasteners can now be measured by ULTRASONIC means (see ULTRASONICS). This involves measuring the initial length of a fastener using ultrasonic equipment. The "acoustic length" of a bolt will almost always differ from the actual length because of minute differences in alloy content, heat treatment, residual stress, etc.-all of which affect the velocity of sound. The change in physical length will be essentially equal to the change in acoustic length, however, as the bolt is tightened or loosened.

LENGTH, EFFECTIVE

(See STIFFNESS OF FASTENER AND JOINT) When controlling .preload by measuring bolt stretch (mechanicaliy

or ultrasonically) it is often useful to be able to predict the relationship between total elongation of the bolt and applied preload. To do this we must estimate the stiffness or elasticity of the fastener. A long bolt, for example, will stretch more under a given preload than will a short

. bolt of the same diameter. For purposes of stretch or stiffness calculations, however, the "length" of the bolt is not its total length, but only that portion of the length which is loaded as the nut is tightened. Threads sticking out beyond the nut, for example, will be free of any load. That portion of the length which contributes to overall stretch is called the effective length.

lf the fastener is fully threaded, its effective length is usually assumed to be equal to the GRIP LENGTH of the joint (LG) plus one-half the combined thickness of both nuts (HN1 + HN2).

If the fastener has a body (is not fully threaded), then we must compute and add the effective length of the body to the effective length of the threaded portion of the fastener.

Effective length of the body is the length of the body within the grip (LBG) plus one-half the height of the head (HH).

LB = LBG + HHI2 The effective length of the threads is the length of the threads with

in the GRIP LENGTH (LSG) plus one-half the height of the nut (HN): LS = LSG +HNI2

The effective length of the fastener (LE) is: LB + LS.

68 Bolting PI;/--- ~ures Reference Manual

FIGURE 19 The effective length of a bolt (LE) is equal to the sum of the effective lengths of the body (LB) and of the threaded region (LS). The body's effective length includes half the thickness of the head. The thread's effective length Includes half the thickness of the nut.

~LB--~

( LE--~)

LOOSENING

(See VIBRATION LOOSENING)

LUBRICANTS

(See Also GALLING; NUT FACTOR)

Bolting Procedures Reference M, 69

General

Lubrication of the fastener threads and the bearing surface of the turned element (the nut or the bolt head) is essential when torque is used to control preload. Two photographs illustrate the importance of proper lubrication.

FIGURE 20 Photograph of galled threads. Lubricant had been used on these threads when first assembled.

Thread Friction

Fig. 20 shows the galled threads of a 3 inch diameter A354 stud which had been repeatedly preloaded to 450,000 lbs by applying a torque of 20,000 ft-lb. The?ge is evidence of the high surface loads and friction forces on the threads.

It is estimated that 40% of the torquing effort is nonnally used in overcoming friction at the thread surfaces. However, when threads gall, 100% of the effort goes into twisting the fastener with no increase in preload.

70 Bolting Pr ures Reference Manual

FIGURE 21 Galled surface of a nut assembled without lubricant.

Bearing Surface Friction

Fig. 21 shows the bearing surface of a 3" diameter nut which was torqued to 20,000 ft-lb WITHOUT lubricant. The surface damage was caused by the high bearing loads and frictional forces which act on this surface during torquing.

Under normal conditions, it is estimated that 50% of torquing effort is used in overcoming friction at the nut bearing surface. When galling occurs, the percentage is much higher.

Nut Factor

An empirical nut factor (K) is used to relate torque to preload in the following short-form equation:

T ~ K D Fp / 12 (See TORQUE, SELECTION OF) The nut factor is primarily determined by the thread lubricant. (see

Table in NUT FACTORS).

Bolting Procedures Reference Mar 71

Note that there are many other factors in addition to lubricants which affect the torque/preload relationship (surface finish, hardness, angularity of parts and purity of lubricant); therefore, it is not surprising to find a large range of reported nut factors. For most assemblies the mean nut factor is adequate for calculating the required torque. If the preload is critical, the nut factor can be measured by means of stretch measurements during assembly (see ASSEMBLY PROCEDURES-Stretch Control).

Again, for emphasis, it is important to remember that the "K" value is an experimentally derived constant. The "K" value has to be redetermined for each new application involving changes such as:

o Change in type of fastener, e.g., from studs to bolts o Change in fastener size, thread type, thread surface such

as machined versus rolled threads, etc.

Coefficient of Friction

It should be noted that the nut factor "K" is not the same as the coeffi· cient of friction, which is just one of the many factors which determine the nut factor.

Selection of Lubricant

Consider the following when selecting a lubricant:

o Compatibility. The lubricant must be compatible with the fastener material and with the contained fluid. Chlorides, fluorides and sulfides are undesirable since they contribute to stress corrosion cracking. Copper-based lubricants can contaminate primary fluids.

o Lubricity. The NUT FACTOR Table illustrates a wide range of nut factors. A lower nut factor is indicative of a more efficient lubricant.

o Temperature. Each lubricant has a recommended service temperature limit. See the manufacturer for information.

72 Bolting· ~dures Reference Manual

Use of Lubricant

o Use only specified or approved lubricants on assemblies. o Apply the lubricant in a consistent manner:

-Lubricate both threads and bearing surfaces. -Avoid over-lubricating by globbing lubricant on the parts,

as this may reduce the efficiency of the lubricant. -Apply a thin uniform coating of lubricant to the parts.

Note that it's also important to keep the supply of lubricant covered when not in USe and to store it in a clean, controlled area.


1. Czajkowski, C. L Testing of Nuclear Grade Lubricants and Their Effect on A540 B24 and A193 B7 Bolting Materials. Brookhaven National Laboratories. Department of Nuclear Energy. New York. March 1984.

2. Rowland, M. C. and T. C. Rose. Tests on Thread Lubricants in Nuclear Reactors. General Electric Co., Nuclear Energy Div., San Jose, California. July 31, 1968.

MATERIALS, SOL T

(See MATERIALS, IDENTIFICATION OF)

MATERIALS, IDENTIFICATION OF

The easiest way to determine the material a bolt is made of is by reading its head marking. A. number of these are tabulated below.

If there is no head marking, or you're still in doubt, you should consider measuring the hardness of the fastener (see HARDNESS OF FASTENER). This can give you a rough estimate of yield strength, at least.

If identification must be absolute, have a qualified metallurgical laboratory test the material for you.

Bolting Procedures Reference Manu, "-\73

FIGURE 22

Specification ASTM A 193

Idotnll1lo:l1tlon G .... QradeM"k Mat.,.IBL

" 8 AISI ':140

'"

'" e AISI 321

Q Al5I 304

" Strain Hardened

e AISI 347

'" Strain Hardened

e Al51 :H6

'" Strain Hardened

8 A15I 321

'" Strain Hardened

S cification AS1'M A 490 ,.

e Alloy Steel, Type Quenched and Tempered 1

9 A.tmospheric Corrosion Type (Weathering) Steel, , Quenched and Tempered

74 Bolting F

FIGURE 22 Continued

:lures Reference Manual

SpecificatiOn ASTM A 325

Type e Medium Carbon Steel,

1 Quenched and Tempered Radial Dashes Option .. l

Type 8 Low Carbon Martensi ttl , Steel, Quenched and Tempered

Type G Atmospheric Corrosion , (Weathering) Steel Quenched and Tempered

Spedf" ti 5 '0' 00 " J 429

Idenlillcalion Grade G,adeMark Male,;al

0 0 , Low or Medium , Carbon Steel

, 0 , 0 Medium Carbon Steel.

Quenched and Tempered

,. , 0 LOW Carbon Martensitlc Steel, Quenched and Tempered

, 0 Medium Carbon Alloy Steel, Quenched and Tempered

, 0 Medium Carbon Alloy steel, Quenched and Tempered

'.1 0 Medium Carbon Alloy or SAE 10":1 Modified Elevated Temperature Drawn Steel

'-' 0 Low Carbon Martens; tic Steel, Quenched and Tempered

FIGURE 22 Continued

po. cation ASTM A

IdentifIcation

Grid, GfadeM.fk

" 0) ", e m e "" e '" e " 0) ,eo e Bet e eo, e

Bolting Procedures Reference Me 75

'" Mat&,ia\

AISI 41 40, 4142 0' 4145

AlSl 4037

AISI 4137

AISI 8740

AISI 4340

AISI 304

AISI 347

AISI 321

AISI 303 or 303Se

76 Bolting Pr,-'-~-- 'ures Reference Manual

FIGURE 22 Continued

Specification ASTM A 320 (Continued)

''" 8 AISJ 316

" (0 AISI 30<1

'" G AISI 3<17

'" Q AlSI 303 or 3035e

"M G A!SI 316

'"' G AISI 321

Specific"tion ASTM A 35<1

" 0 Alloy Steel,

0 Q~'enched "nd Tempered

"

Specification ASTM A <149

o Medium C"rbon Steel, Quenched and Tempered

Bolting Procedures Reference Mr 77

MATERIALS, PROPERTIES OF

(See property of interest under: ULTIMATE STRENGTH; YIELD STRENGTH; COEFFICIENT OF EXPANSION; MODULUS OF ELASTICITY, etc.)

MODULUS OF ELASTICITY

Also called Young's Modulus, the modulus of elasticity is the ratio between stress in a body and the resulting strain. The chart below gives modulii for various bolting materials at temperatures between 70 deg.-800 deg. F.

78 Bolting Pro(-"-" ,res Reference Manual

FIGURE 23 Modulus of elasticity of various bolting materials as a function of bolt material.

"

"

"~"----~""O--~30~O----'~OO'---~--~~--~~--~ HMPERATURE 'f

\ = Carbon Steel (A307,A36, Carbon =(0.30%) ~ = Carbon Steel (Carbon) = 0.30% ~ = Carbon Molybdenum Steel ) = Nickel Steel (88, 88M) : := Chrome Molybdenum Steel (1/2 to 2 Cr) : = Chrome Molybdenum Steel (2 1/4 to 3 Cr) 3 = Chrome Molybdenum Steel (5 to 9 Cr) , = Straight Chromium Steel

= Austenitic, High Alloy Steel (A453)

leference: Table 1-6.0 of Appendix I, Section III, Division I ASME Boiler and Pressure lessel Code, Nuclear Power Plant Components.

Bolting Procedures Reference 1\ / ~·\al 79

NUT FACTOR

(See Also LUBRICANTS) An experimentally determined factor which sums up the relation

ship between the torque applied to the nut and the achieved preload. The nut factor is defined by the following equation:

K = 12T / (0 Fp) Where: T = Torque (Ft-Lbs) .

o = Nominal diameter (In) Fp = Achieved preload (Lbs)

The relationship between torque and preload is affected by a large number of variables (see PRELOAD, SELECTION OF LUBRICANTS). As a resuit, the experimental constant we can the nut factor is subject to wide variation, depending upon the specific conditions under which it was measured. Each experiment tends to result in a different number.

The table below shows some of the numbers which have been reported in the past. Since the LUBRICANT is often the dominating variable, the data is reported for various lubricants.

If you need maximum preload accuracy, however, you should recognize that many other variables-such as bolt diameter, bolt mater.ial, tightening speed, thread fit and even operator skill can affect the torquepreload relationship. You're well advised to determine the actual nut factor on your application by a test of your own.

If accuracy is less important, you can use the mean values reported below:

80 Bolting Pr/ Ires Reference Manual

TABLE G Nul Faclors (K)

Min. Max. Lube Reported Mean Reported

As-Received 0.158 0.2 0.267 Alloy or Mild Steel Fasteners

As-Received 0.3 Stainless Steel Fasteners

Cadmium Plate (Dry) 0.106 0.2 0.328 Copper Based Anti-Seize 0.08 0.132 0.23 Cadmium Plate (Waxed) 0.17 0.187 0.198 Fel-Pro C54 0.08 0.132 0.23 Fel-Pro C-670 0.08 0.095 0.15 Fel·Pro N 5000 (Paste) 0.13 0.1~ 0.27 Machine Oil 0.10 0.21 0.225 Moly Paste or Grease 0.10 0.13 0.18 Never-Seeze (Paste) 0.11 0.17 0.21 Neolube 0.14 0.18 0.20 Phos·Oil 0.15 0.19 0.23 Solid Film PTFE 0.09 0.12 0.16 Zinc Plate (Waxed) 0.071 0.288 0.52 Zinc Plate (Dry) 0.075 0.295 0.53

NOTE: It is important to remember that the "K" value is an experimentally derived constant.

PRELOAD

(See PRELOAD, INITIAL; PRELOAD, RESIDUAL) Preload is the tension force developed in the fastener when it is tight

ened against the joint.

Bolting Procedures Reference r.. lal 81

PRELOAD, CONTROL OF

(See ASSEMBLY PROCEDURES; PRELOAD, SELECTION OF; TORQUE-SELECTION OF)

PRELOAD, INITIAL

(See PRELOAD; PRELOAD, RESIDUAL) Initial preload is the tension created in a fastener when it is first tight

ened, before the wrench has been disengaged from the nut.

PRELOAD, LOSS OF

Causes

For a number of reasons, the preload in a bolt is often less than you expect it to be; a fact you discover when you check preload by applying breakaway or restarting torque, or when you disassemble the joint. This low preload can be caused by a number of factors. For example:

Low Initial Preload

Even though you've used a proper bolt-up procedure, have wellcalibrated tools, and used a clean lubricant, a good percentage of the bolts will have less preload than assumed. For factors which affect the preload achieved for a given assembly torque, see ACCURACY and PRELOAD-SELECTION OF.

82 Bolting Pr( 'ues Reference Manual

Embedment

High spots on the contact surfaces of threaded fasteners, especially new ones, tend to creep and flow when initially loaded. Preload is lost as the parts settle in together. The loss can be 5-10% in the first few minutes after tightening. See EMBEDMENT for further details.

Cyclic Load Embedment

Joints subjected to cyclic loads, especially large loads, will embed, and therefore relax, more than joints under static loads. If external loads approximate the yield strength of the bolt, preload losses of 25% or even 50% can occur.

Gasket Creep

Gaskets must be partially plastic to function. As a result, they too will creep after initial loading. Preload loss at room temperature can be 2-5% or so; and will occur in 10-20 minutes after initial loading. See GASKETS, CREEP OF.

Thermal Effects

A jOint subjected to a change in temperature can lose preload for several reasons. Differential expansion between bolts and joint members can increase stresses in all parts and, therefore, increase embedment or ,asket creep. Or the bolt can "expand away" from the jOint. The gasket can be compressed beyond the original compression and, due to hys:eresis, won't fully recover when the temperature change is reversed. rhe creep of bolts and gasket can be promoted by high temperature • Ione (even if stresses aren't increased). A process called stress relax.tion can cause loss of preload over a period of time if temperatures Ife high enough. See THERMAL EFFECTS and STRESS RELAXATION or ways to estimate and compensate for these losses.

--------------------'- --Bolting Procedures Reference 1'\ \1 83

Self.Loosenlng

Vibration, flexing of the joint, cyclic shear loads, thermal cycles and other factors can cause whole or partial selfloosening of a fastener. See VIBRATION LOOSENING for a discussion of this phenomenon and for ways to combat the loss.

Elastic Interactions

Tightening one bolt will often allow a previously tightened neighbor to relax a bit. See ELASTIC INTERACTIONS for details.

How to Minimize Preload Loss

EMBEDMENT loss, early gasket creep and, to some extent, ELASTIC INTERACTIONS, can be compensated for by retightening the bolts after giving them some time to relax. Since bolts will interact during any tightening pass merely retightening them won't eliminate interaction loss, but it can red1.lce it. A final torquing pass in a reverse sequence can be helpful, for example.

Increasing initial preloads are a common way to compensate for relaxation of various kinds. They still relax, but residual preloads are higher because they relax from a higher initial preload. Higher preload also helps resist selfloosening as discussed under VIBRATION LOOSENING.

Embedment and creep can be compensated for by tightening the bolts to rated preload, waiting a bit, then loosening and retightening them. This process works the threads and joints in together. Some people call this" artificial aging" of the joint. One aircraft engine manufacturer tightens and loosens a main spindle nut half a dozen times (to a fraction of final preload) before final tightening .

Making the baIts more flexible helps fight all types of relaxation. They'll still embed or creep or etc. by the same amount, but that change in elongation means less loss in preload. See STIFFNESS OF BOLT AND JOINT for ways to increase flexibility.

Elastic interactions can be overcome by using DATUM RODS, strain gaged load cells, or ULTRASONICS to measure residual preloads after the joint has been tightened, then retightening by the indicated amount. See STRETCH CONTROL for more information.

If you have thermal problems, you'll want to re-consider the materials you're using and their COEFFICIENTS OF EXPANSION. Higher initial preloads and more flexible bolts can also help (see THERMAL EFFECTS).

84 Bolting, 'dutes Reference Manual

PRELOAD, RESIDUAL

(See PRELOAD; PRELOAD, INITIAL; PRELOAD, LOSS OF) The remaining tension in a previously tightened fastener after such

things as ELASTIC INTERACTIONS, EMBEDMENT, gasket creep, etc. have allowed the fastener to relax a little from its initially tightened condition.

PRELOAD, SELECTION OF

General

The main purpose of the fastener is to clamp two or more jOint members together. The behavior of the joint-whether it leaks or slips or shakes apart, etc-depends upon the amount of clamping force on the joint. Although service loads and other things affect it, we determine the amount of that clamping force when we tighten the fasteners; when we develop the initial tension or "preload" in them. Picking the correct preload and achieving it in practice are very important steps.

In spite of this fact, we'll pick most preloads indirectly by picking a reasonable assembly torque. Since most joints are overdesigned and can stand wide variations in preload, this is acceptable. In more important joints, however, we must carefully pick a correct preload before going on to convert our selection to torque or tUrn or stretch Or some other control variable.

In this section we'll consider a variety of ways to select preload itself, starting with simpler ways and proceeding to more complex ways. In general, you should always use the simpler ways unless:

o you've used the simpler ways on this joint in the past and results have been poor or

[J you have no prior experience with this joint, but are concerned about potential joint problems.

Bolting Procedures Reference M 85

Preload Selection Options

Here's a "menu" of the preload selection options discussed belm~. In general, as you move from one Level to a higher one, there will

be an improvement in the" accuracy" of your selection of preload. thiS means that the preload selected will be more carefully tailored to the specific problems you've experienced andlor are concerned about.

Levell

Level 2

Level 3

Beyond Level 3

Repeat your previous procedures if your prior experience with this joint has beer: acceptable or get a re<;ommendation from the vendor if you have no prior experience. . Select a torque from simple tables (the :ollecllv~ experience of others) if you have no pnor expenence with this joint, and are not overly concerned about potential problems. The tables we've provided below assume that the fasteners are not lubed, that they operate at room temperature, and that an average bolt stress of 50% of yield is acceptable. If prior experience coul~ ~e better, or you have so~e concern about possible Jomt problems, you should. -determine the yield strength of the bolts at the

service temperature -use a table to select an appropriate percentage of

yield for your application -make a simple calculation on a pocket calculator to

compute a preload -select an appropriate nut factor from a table, to ac

count for the lubricity of the bolts -use the calculator again to compute torque.

Beyond Level 3 you will make increasingly complicated decisions and calculations to compute preload

d if desired to convert it to an assembly torque. an " . You will also have to pay more attenllon t~ opera-tional problems. A complete treatme';"!, which you will need only if all else fails (or if failure must be avoided at all costs) would consider such factors as:

86 Bolting l ... ,dures Reference Manual

-Loads on the bolts: Mechanical (weight) Pressure Thermal Seismic Fatigue Vibratory Shear Torsional Misalignment Bending Prying Eccentric

-Upper limits on bolt tension: Thread stripping limits Str~ss corrosion cracking limits FalIgue limits Joint strength Gasket stress limits

-Assembly variables: Scatter in lubricity Tool accuracy Elastic interactions between bolts

-Response of the joint: Relaxation effects Flange rotation Influence of jOint/bolt stiffness ratio

The various procedural levels we've established for thO I described below for your guidance: IS manua are

Level 1 Procedures-Prior Experience Or Vendor Advice

It is difficult and expen' t . SIve 0 get accurate control of preload in the field. As a :esult, most joints have been overdesigned to tolerate wide ;~~lalI~ns ill ~reload. In most applications, furthermore, bolt or joint al Ufe ~s a nUIsance, but not a real problem. So we can he relativel

casual In our selection of preload. Y In facbtlwe can, and usually do, select it indirectly by selecting a

reaSOna e torque to use for assembly. . Whet~er Or not we hav~ selected preload by selecting a tor ue, our

best gUlde to success IS pnor experience. The reason? a large ~umber

Bolting Procedures Reference M, 87

of variables affect the preload in a bolted joint; we can never hope to know or control them aU and so must rely on experimentation to predict final results. The closer the "experiments" are to the actual job, the more apt they are to predict results correctly. So actual experience is our best guide. If your past experience with this joint is acceptable, don't change anything!

If you have no past experience to guide you, ask the vendor who supplied the equipment for a recommendation. How much torque does he suggest you use? This information may be found in the manuals provided with the equipment, or on his drawings. Or you may have to give him a call. But checking with the vendor is a smart move at any level of preload selection.

Level 2 Procedure-Simple Torque Tables

If you have no prior experience with this joint, and you're not particularly concerned about failure, then it's safe for you to use the collective experience of others, as summarized in torque tables.

The Simplest table (Table H) takes into account only two variables:

o The size of the fastener. o The material it's made from.

The people who designed the table assumed the following:

o The service temperature of the fasteners will be room temperature (70 deg. F or 20 deg. C).

o The fasteners will be unlubricated. o It is acceptable to tighten them to 50% of yield.

EXAMPLE: Determine the torque to be used on a 1 1/4-8xlO, A193 B7 bolt. The table suggests that we use a torque of 959 ft-lbs.

If your prior experience with Level 2 has been poor or you're too concerned about potential failure to rely on a simple table or the fasteners operate a a temperature above 100 deg.F, then you should move on to Level 3.

E:;:. ~. 88 Bolting o' " edures Reference Manual T~

Bolting Procedures Reference M 'I 89 9:'-TABLE H ~!:!i!J Torque Chart for 50% of Yield Using a 0.2 Nut I . TABLE H Factor and Room Temp Yield Strengths, UNC Thread T

Continued Series

MATERIAL TYPE Material Type T (See Legend) G H J K L (See Legend) A B C 0 E F Yield [ksiJ 105 109 120 130 140 150 Yield [ksiJ 30 36 81 85

p!s~ Oia. 95 100 I Dia. 4 7 7 8 9 9 10

~r 6 13 14 15 16 17 19 4 2 2 5 6 6 7 In· 8 24 25 27 30 32 34 6 4 5 10 11 12 12 Lbs 10 35 36 40 43 46 50 In· 8 7 8 19 20 22 ~.~ 12 55 57 63 68 73 78 Lbs 23 10 10 12 27 28 b 12 31 33 1/4 7 7 8 9 9 10 16 19 42 44 50 52

~I 5/16 14 15 16 18 19 20 1/4 2 2 5 6 6 7 3/8 25 26 29 31 34 36 5/16 4 5 11 12 13 14 7/16 41 42 46 50 54 58 3/8 7 9 20 21 23 24 ~~ 1/2 62 64 71 77 83 89 7/16 12 14 31 33 37 39 ~I 9/16 89 93 102 111 119 128 1/2 18 21 48 50 56 59 ~~ 5/8 123 128 141 153 165 176 9/16 26 31 69 73 81 85 3/4 219 227 250 271 292 313 5/8 35 42 95 100 112 118 -. 7/8 354 367 404 438 472 505 3/4 63 75 169 177 198 209 ~.~ Ft- 1 530 550 606 656 707 757 7/8 101 121 273 286 320 337 -I~ Lbs 1 1/8 751 780 858 930 1001 1073 FI- 1 151 182 409 429 480 1 1/4 954 1100 1211 1312 1413 1514 505 Lbs 1 1/8 214 257 579 608 679 ~r 1 3/8 1390 1442 1588 1720 1853 1985 715 1 1/4 303 363 817 858 959 1009 1 1/2 1844 1914 2107 2283 2459 2634 1 3/8 397 476 1072 1125 1257 1323 ~.':3 1 3/4 2909 3020 3325 3602 3879 4156 1 1/2 527 632 1422 1493 1668 1756 2 4375 4542 5000 5416 5833 6250 1 3/4 831 997 2244 2355 2632 2771 ~~ 21/4 6398 6642 7312 7921 8531 9141 2 1250 1500 3375 3542 3958 4167 2 1/2 8750 9083 10000 10833 11667 12500 2 1/4 1828 2194 4936 5180 5789 6094 -,~ 23/4 11863 12315 13557 14687 15817 16947 21/2 2500 3000 6750 7083 7917 8333 3 15671 16268 17910 19402 20895 22387 23/4 3389 4067 9151 9603 10733 11298 ~.~ 3 1/4 20191 20960 23075 24998 26921 28844 3 4477 5373 12089 12686 14179 14925

~b 3 1/2 25511 26482 29155 31585 34014 36444 3 1/4 5769 6922 15576 16345 18268 19229 3 1/2 7289 8746 19680 20651 23081 24296 LEGEND:

LEGEND: I Material types included in each column: ~.:::!I G-ASTM A540 821 CL 5 TO 2" OIA, 822 CL 5 TO 2" OIA, 823 CL 5 TO 6" OIA, Material types included in each column: .- 824 CL 5 UP TO 6" OIA; A193 87 UP TO 2.5" OIA, 816 UP TO 2.5" OIA; A437

A-ASTM A193 & A320 GR 88M, B8T, 88C, 88; SAE 304, 316; AISI 1020 848; A320 L7 UP TO 2.5" OIA, L43 UP. TO 4" OIA 8-SAE J429 GR 2 FOR OIA OVER 3/4 TO 1 1/2, Gl; AIS11038; ASTM A307

~ H-ASTM A354 GR 8C 1/2 TO 2.5" OIA C-ASTM A325 TYPE 1,2&3; A449 OVER 1" TO 1 1/2" OIA SAE J429 GR 5 OVER 1- ASTM A540 821 CL 4 UP TO 6" OIA, 822 CL 4 UP TO 4" OIA, 823 CL4 UP TO 1" TO 11/2" DIA '

~-:i' 9.5" DIA, 824 CL '4 UP TO 9.5" O-ASTM A193 B16 FOR OIA OVER 4" J- ASTM A540 821 CL 3 UP TO 611 Of A, 822 CL 3 UP TO 411 D1A, 823 CL 3 UP E-ASTM A193 87 & 816 OVER 2 1/2" TO 4" OIA; AISI 4340 • TO 9.5" OIA, B24 CL 3 UP TO 9.5"; A354 GR 80 UP TO 2.5" OIA; A490 TYPE F- ASTM A540 821 CL 5 OVER 2" TO 8" OIA, 822 CL 5 OVER 2" TO 4" OIA 823 !i!:':-:i' 1; SAE J429 GR 8 & 8.1 1/4 TO 1 1/2 OIA, GR 8.2 1/4 TO 1" OIA.

CL 5 OVER 6" TO 9" OIA, B24 CL 5 OVER 6" TO 9" OIA; SAE J429 GR 4'

~!~ K- ASTM A540 821 CL 2 UP TO 4" OIA, 822 CL 2 UP TO 3" OIA, B23 CL 2 UP

TO 9.5" DIA, 824 CL 2 UP TO 9,5/1 DIA; AISI 4140 L- ASTM A540 821 CL 1 UP TO 4" OIA, B22 CL 1 UP TO 1.5" OIA 823 CL 1 UP

I TO 8" OIA, 824 CL 1 UP TO 8" OIA

~~ ~.:iI

I ~i=:

90 Bolting:/,'0 -, edures Reference Manual

Level 3 Procedure-A Simple Calculation

In selecting a preload at Level 3 we'll now deal with four variables:

o The size of the fastener. o The material it's made from. o The actual service temperature of the bolts. o The percentage of yield suggested by the application.

If :ve wish to convert the resulting preload to torque, we'll also conSIder:

o The actual lubricant used on the fasteners.

Since this is still "only Level 3", however, we'll get the input data we need from tables.

Th~ cal:ul~tions can be made on any pocket calculator. Only multiphcahon IS mvolved.

(Note that we could also construct Torque Tables for Level 3. With five variables to deal with, however, the tables become very long and cumbersome. We think you'll find the calculator approach more convenient.)

Level 3 Worksheets have been provided with this manual. We suggest that you duplicate and use them.

The calculations you'll make are based on the following equaion: T ~ I<: D M Sy As I 12

Where: T ~ Torque (ft-Ibs) D ~ Nominal diam .. (in) M ~ Percentage of yield (as a decimal) K ~ Nut factor .

Sy ~ Yield strength (psi) As ~ Tensile stress area (in2)

All of the data required to solve this equation will be found in various tables in this manual. References are given in the manual and on Worksheets.

. ;::he first decision you'll h~ve to make when using the Worksheet ~s What percentage of Yield IS correct for this application?" Since this IS a Level 3 procedure, the decision is not a critical one and can be made by considering f'common practice". Examples are' given at the beginning of the worksheet.

Bolting Procedures Reference tv '\ ,I 91

Level 4-Taking Problems and Loads Into Account

If Level 3 hasn't solved your problem or hasn't provided enough assurance, you should start to take a closer look at the loads on this specific joint, the problems you've had with it and the procedures used to assemble it as discussed below.

Operational Considerations If you've reached this level, but are still having problems, you should probably pay additional attention to assembly practices. Your problems may be caused not by an incorrect preload or torque specification, but by the fact that you're not really achieving that specified preload in the joint. Training the operators more carefully, supervising them as they assemble the joint, can often solve chronic problems more readily than changing the torque.

Variation or scatter in the amount of preload achieved for a given torque is another common problem which prevents a good choice of preload or torque from solving or preventing a joint problem. A large number of operational, material and quality factors contribute to this scatter. The results can be summed up by assuming a scatter in the nut factor, K, which in previous levels, we've treated as a constant for a given lubricant. You'll find scatter data along with nominal values in the Table under NUT FACTORS. Here's how to use that data to estimate the scatter in preload:

Nominal Preload (in terms of applied torque) Fp~12T/KD

Maximum Preload Fp ~ 12 TI KMinD

Minimum Preload Fp 12 TI KMaxD

Where: T ~ Applied Torque (ft-lbs) D ~ Nominal Diameter (inches) K ~ Nominal NUT FACTOR

Fp ~ Preload (lbs) KMin ~ Minimum Reported Nut Factor KMax ~ Maximum Reported Nut Factor

Does the computed scatter in preload explain your continuing problems?

Another way of looking at this: An actual nut factor 25% higher than the nominal would have the

same effect as applying 25% less torque and preload than you specified.

92 Bolting V " ,dures Reference Manual

Using Past Experience Past experience is still your best guide. It can often help you decide how to modify your previous preload (or torque) selections to solve a problem. For example, if you've had any of the following problems, you should probably INCREASE preload (torque) the next time you assemble that jOint. Typical "low preload" failure modes include:

o Leaks o Vibration Loosening o Joint Slip, Fretting o Fatigue

Before you increase preload (torque), the design basis of the components should be checked to ensure that nothing is overstressed. Look at flange distortion, gasket stress, fastener strength and thread stripping. ,

If you need help identifying a failure mode, see those sections of the manual which discuss the various types, LEAKS, FATIGUE, etc.

How much increase should you try? That will require some judgement. If you think that the previous torque almost worked, you should perhaps try 10% more torque this time. If you were way off, try 20% or 30%. Of course, if your previous choice was over 60-70% of yield, you can't try 30% more without also doing something about the way m whIch you trghten them. You must improve the accuracy of the tightening process to avoid taking the bolts past yield. (See ACCURACY.)

You'll want to DECREASE preload (torque) a little or a lot if you've had these problems;

o Stress Corrosion Cracking o Thread Stripping o Damaged Joint Members (or Gaskets) o Leaks Caused by Excessive Flange Rotation

Decreasing preload can be a risky thing to do, however, unless this is a relatively unimportant jOint and/or is subjected only to modest, static loads. If you think that a decrease might lead to some of the low preload problems listed earlier (leaks, vibration loosening, etc.), then you should probably consult a deSigner before making the change.

You should also consult a designer if you've already had both high and low preload problems with this jOint (for example, stress corrosion cracking and leaks).

Theoretical Considerations Depending upon the nature of your problem or concern, there are some relatively simple calculations you can make at this level to improve your understanding of the loads on

E="-r __

~

~ ~q!1

~ ~ ~~

~ ~

Bolting Procedures Reference M "~--'\J 93

this joint and therefore to improve your selection of preload or torque. Note that in previous levels we have considered only the bolt-its

size, strength, temperature, lubricity-in picking a preload or torque. In Level 4 we'll start to consider certain aspects of the joint as well.

LEAKS (Effect of Pressure Loads) A gasketed joint will leak unless two conditions are met:

o The gasket must be initially seated with sufficient force (usually computed as a seating stress).

o After the system has been pressurized, there must still be sufficient residual stress on the gasket to resist the internal pressure in the system (computed as a multiple of contained pressure, P).

Duplicate and use the GASKET WORKSHEET provided with this manual and see GASKETS STRESS for help in computing gasket stress and the effects of pressure loads.

Thermal Stress Changing the temperature of a previously tightened joint can either increase or decrease bolt and joint stress levels-or leave them unchanged -depending upon the thermal COEFFICIENTS OF EXPANSION of bolts and the joint.

If the joint has a larger coefficient of expansion than the bolts, and/or becomes hotter than the bolts, it will expand more than the bolt. Since it is held between bolt head and nut (or two nuts), this expansion will both increase tension in the bolt and the clamping force between joint members.

If the bolts have a higher coefficient, or are hotter than the joint, then the tension in the bolts, and the clamping force on the joint, will be reduced.

See THERMAL STRESS for a procedure and worksheet for estimating the effects of a change in temperature on preload and clamping force.

Beyond Level 4

By this point we've considered all the relatively simple and common factors used to select a target preload (and torque). We worried about the size and estimated lubricity of the fastener and considered tool accuracy and thermal effects. And we've taken external loads (pressure, for example) into account-at least in a simplistic way. If the joint still misbehaves, however, there are adelitional factors which can be considered. These are complex and must be dealt with by a qualified engineer.

94 Bolting/'··· -:edures Reference Manual

A full discussion would be beyond the scope of this manual, but the engineer can consider things like these in refining the choice of target preload (see the sections referred to for further definitions):

o FLANGE ROTATION, which can sometimes result in a reduction of GASKET STRESS as the bolt tension is increased.

o Time-based THERMAL EFFECTS such as STRESS RELAXATION, and CREEP.

o The actual way in which an external load, such as a pressure load, is absorbed by the bolts and joint members.

o Optimum gasket stress. The Subcommittee on Bolted Flanged Connections of the PYRC has developed a new way to select "m" and "y" gasket factors. The procedure is more complex than selecting m and y from a table, but is firmly grounded on experimental data. It can help solve chronic leak problems. (See GASKET STRESS for references.)

o Depending upon the nature of the problems you've experienced with the joint, the designer may also have to consider-and/or consult an expert on -VIBRATION LOOSENING, FATIGUE or STRESS CORROSION CRACKING. A finite element analysis of thermal and stress patterns may be required. Offset PRYING or ECCENTRIC load may be contributing to the problem. Perhaps relaxation effects are creating PRELOAD LOSS which should be compensated for in the selection of initial preload.

PRESSURE BOUNDARIES

(See ASSEMBLY PROCEDURES, GENERAL-All Joints; Leaks)

__________________ r·, Bolting Procedures Reference M ;:1 95

PREVAILING TORQUE The torque required to run a virbation resistant nut down to the jOint

surface when some If obstruction", such as a nylon colalr or distorted thread form, is present in either male or female threads. Prevailing torque musthe added to the clamping torque required when selecting the torque to be applied to this type of vibration resistant fastener.

FIGURE 24 A vibration resistant nut with a nylon "collar".

nylon

PROCEDURES, ASSEMBLYI DISASSEMBLY

(See ASSEMBLY PROCEDURES, GENERAL-All Joints; DISASSEMBLY PROCEDURES)

96 Bolting r dures Reference Manual

PROOF LOAD

The maximum safe tensile load which can be applied to a fastener without creating any permanent deformation. Proof load is usually about 90% of the minimum yield strength.

PRYING

External tension loads on a fastener are magnified by a sort of lever action called prying when the line of action of the external load does not lie along the axis of the fastener. Prying almost always exists, but is not a problem unless joint members are quite flexible and/or the line of action of the load is a considerable lever distance from the bolt axes.

FIGURE 25 An Illustration of prying action.

Axis of Bolt

I r

References, General The following documents and texts have helped us prepare this manual-and can give you additional information.


1. Bickford, J. H., An Introduction to the Design and Behavior of Bolted Joints, Marcel Dekker Inc., New York, 1981.

2. Fisher, J. W. and J.H.A. Struik, Guide to Design Criteria for Bolted and Riveted Joints. John Wiley and Sons, New York, 1974.

3. Yahr, G.T., Preloading of Bolted Connections in Nuclear Reactor Component Supports. NUREG/CR-3853, ORNL-6093, 1984.

4. Product Engineering Report No. 4718-Fastener Seminar. SPS Technologies, Jenkintown, Pennsylvania, 3rd printing. 1980.

5. Davis, G. W., A Training Program in Quality Assurance Aspects of Fasteners in Nuclear Power Plants. Prepared by Center for Nuclear Studies of Memphis State University for the Office of Power, Tennessee Valley Authority, July 1982.

6. Utility Recommendations and Guidelines for the Purchase Specification and Receipt/Preinstallation Inspection Requirements for ASME Section III, AISC, ANSI! ASME B31.1 and ANSI B31.5 Bolts and Threaded Fasteners. Prepared by the Joint AIF/MPC Task Group on Bolting Requirements. Pub!. by EPR!. May 1985.

RELAXATION, FASTENER

(See PRELOAD, LOSS OF)

ROTATION OF FLANGE

(See FLANGE ROTATION)

98 Bolting p;/ - 'dures Reference Manual

SEALANTS

A number of potential concerns have been identified regarding nuclear plant leak sealing (i.e., increased loads on components, added weight on piping systems, sealant intrusion into the process line, increased corrosion potential. There are two primary considerations: (1) maintenance of pressure boundary integrity both during and after the sealant injection and (2) the effect on system and component function and operability due to the sealant injection. Each of these considerations should impose restrictions and/or check points on proposed leak sealing operations, which will depend on the safety significance of the components to be sealed.

The following general recommendations are made:

o Do not perform leak sealing operations on the reactor coolant pressure boundary (ASME Section III, Class 1, or equivalent classifications without concurrence from the responsible design organization.

o Do not perform leak sealing operations on active components (any Code class) without prior concurrence from the responsible design organization.

o Do not inject valve packing sealant on active or non~active dynamic valves without prior concurrence from the responsible design organization.

o Consider all applications of leak sealants to be temporary solutions (i.e., repair or replace the leaking components at the next available system shutdown). For repairs, remove all sealant and restore the component to its original configuration or approved alternate.

o In order to ensure adequate leak sealing operations, it is recommended that utilities develop a controlled standard practices document or some other specification which is readily available at the site level, which clearly defines restrictions and check points on all leak sealing operations to assure that design integrity is maintained.

The following recommended guldelines apply to leak sealing operations in general:

o Provide Certified chemical analyses for each batch of sealant. This is a very important consideration, since the sealant will typically be in intimate contact with flange bolting, interior valve parts, etc., and may even find its way into

Bolting Procedures Reference Ma' 99

the process fluid (see below). Test and document each sealant batch or lot to esablish chemistry compliance.

o Monitor sealant cavity pressure (generally conservatively estimated by the sealant injection pressure) closely during the injection operation. Establish overpressure check points prior to injection.

o Prior to injection, determine the anticipated volume of sealant needed to fill the volume of the sealant cavity (enclosure). Establish a check point based on this volume to limit the injection of sealant. This will provide reasonable assurance that excess sealant is not forced into the process fluid.

o The added mass of repair fixtures may compromise the seismic qualification of any piping systems employing such repairs (nuclear plants). Consider seismic requirements prior to repair.

o Consider the design basis for the fixtures employed because these sometimes rather massive structures do become a new pressure boundary once the repair is made. Further, if the repair fixture captures injected sealant between flange faces, the additional loading on the flange studs due to injection pressure (above system pressure), possible increase in effective sealing diameter and mechanical sealing operations must be considered.

References The following document has helped us prepare this section-and can give you additional information.

1. Merrick, E.A., A. Rivers, J. Bickford and T. Marston. Prevention of Bolting Degradation or Failure in Pressure Boundary and Support Applications. Paper given at SMIRT-8 Conference, September 1985.

STIFFNESS OF FASTENER AND JOINT The bolted joint can be modeled as a system of springs: a heavy "joint

spring" and, numerous lighter, IIbolt springs". The way the system absorbs external pressure, weight, seismic and other loads depends upon the relative stiffnesses of these various springs, so stiffness is

1 00 Boltinv-~- Jcedures Reference Manual

of considerable interest to the designer. It is often desirable to reduce the bolt stiffness to reduce thermal effects

and improve fatigue life. To do this you can:

o Turn down a portion of the body and/or threads within the grip length.

o Drill a hole down the center of the bolt. o Use collars or belleville springs on one or both sides of the

jOint so that you can use a longer bolt of the same diameter.

FIGURE 26 Coilars are added to • bolt to Increase Its flexibility without changing Its diameter.

,----'-----L--~ COLLARS

~~

Making the joint stiffer is also helpful, but your options are more limited here. Sometimes, however, you can do this by switching to a stiffer gasket or by providing stops on a gasketed joint. (You'd better consider the possible LEAK implications before making such changes, however.)

Using a hard washer can also increase joint stiffness (by "involving" more joint material) without increasing the dangers of a leak.

~

~ !!";"E!

i

~'7'~

Bolting Procedures Reference Mal 101

STRENGTH OF BOLTING MATERIALS

(See STRENGTH OF FASTENERS) We're usually interested in the static strength of bolting materials

strength defined as ULTIMATE STRENGTH or YIELD STRENGTH. If operating temperatures are below 150 deg. F or so, furthermore, we're only concerned about the room temperature values. ULTIMATE STRENGTHS are given in Table D in the discussion of HARDNESS OF FASTENERS. Table I under STRENGTH OF FASTENERS-STATIC also includes information on ultimate strengths.

The strength of bolting materials will decrease as operating temperatures rise. You must take this fact into account when selecting PRELOAD or assembly TORQUE. You must also consider strength at operating temperature when dealing with certain types of FAILURE OF FASTENER. See YIELD STRENGTH for a chart showing the relationship between strength and operating temperature.

Some bolting problems can be solved by switching to stronger materials having higher yield and ultimate strengths. Increasing static strength, however, is associated with an increase in HARDNESS; and this often means a reduction in ductility. The fasteners become more brittle and more susceptible to STRESS CORROSION CRACKING failure. You should consult a designer and metallurgist before switching materials.

STRENGTH OF FASTENERS-GENERAL

Fasteners can fail in many different ways. These various "strength'! problems are covered under:

o FATIGUE FAILURE (breaks under dynamic or cyclic loads) o STRENGTH OF FASTENERS-Static (the body breaks under

steady state or slowly increasing loads; for example, as they're being torqued)

o STRESS CORROSION CRACKING (sudden failure under static loads and corrosive attack)

o THREAD STRIPPING

----. -_._---------------102 Bolting Procedures Reference Manual

STRENGTH OF FASTENERS-STATIC

If fasteners break under steady or slowly changing loads (for example, as they are being tightened), the forces on the fastener have exceeded its static strength. (See also STRESS CORROSION CRACKING.)If they fail under dynamic or cyclic loads, it's probably a FATIGUE FAILURE.

Static strength of the body is determined primarily by two things: the size of the fastener, and the material it's made from. (The strength is also affected by a number of secondary factors which we'll discuss later.)

Influence of Size

Doubling the diameter of a fastener will increase its static strength to (roughly) four times the original strength. Equations for strength are given under Static Strength Equations, below.

Materlai Selection

Since we can't often change the diameter of fasteners in field situations, we usually select a new material if we wish to increase the strength. If static strength is the only consideration, we merely use a material with a higher strength. You'll find more information on the strength properties of various fastener materials under YIELD STRENGTH. Here's a quick summary of some of the most common choices:

TABLE I

MATERIAL

Low Carbon Steel

Ferritic Stainless Steel Austenitic $.8. (Solution Annealed) Austenitic S. S. (Cold Worked) Med. Carbon Steel (Heat Treated) Austenitic 8.S. (Strain Hardened)

ROOM TEMPERATURE ULTIMATE STRENGTH (IN KSI) EXAMPLES

60 AISI 1021. ASTM A307 GR A. SAE J429 GR 1 & GR 2

70 430. 430F 75 303,304,316,321, ASTM

A193 Cll BS 90 303, 304, 316, 321

120 AIS11030, ASTM A449

125 303,304,316,321, ASTM A193 BS

TABLE I Continued

K-Monel Low Alloy Steel

Titanium Martensitic 8.S. (Heat Treated) Super Alloys

Bolting Procedures Reference' Manuai 103

130 125

150 150

260

K-Monel AISI 4140, 4340, ASTM A193 B7, B16 Ti·6AI4V 410,416,431, ASTM A354 GR 80 H-11, Incone! 718, A·286

The list is for ultimate strengths. We're often more interested in YIELD STRENGTHS-which usually run 65-80% of the ultimate strength for different materials.

Note also in the Chart under YIELD STRENGTH that the strength of these materials decreases as service temperature increases.

The static strength of a bolt is directly proportional to material strength, Increase the ultimate strength by 10% and you'll increase the strength of the bolt by roughly the same amount. If strength is your only concern, that's all there is to it.

But a word of caution. As material strength increases, the material becomes more susceptible to STRESS CORROSION CRACKING.

Secondary Factors

As mentioned earlier, bolt size and material are the main factors which determine its strength. In some situations you may also wish to consider the following "secondary" factors.

Fine pitch fasteners have greater body strength (but less stripping strength) than coarse pitch ones.

In a bolt which is not fully threaded, the tensile strength increases if the nut is tightened to a point near thread run-out (reducing the number of threads within the GRIP LENGTH). This can reduce both THREAD STRIPPING STRENGTH and FATIGUE strength, however,

A well lubricated fastener can be tightened to a higher tensile load before breaking than an unlubricated fastener (because it sees less torsional stress).

Under combined shear and tension loads, long bolts will often support higher loads than short ones because the bolts are more flexible and bend more easily.

A bolt made of material that can be through-hardened, like AISI 4340, is stronger in large sizes than a similar bolt made of an alloy which can't be throughhardened, like AISI 4140.

104 Bolting-r'focedures Reference Manual

Under shear loads, a bolt is stronger if the shear plane is through the body of the boItrather than through the threads. Multiple shear planes also increase shear strength.

Static Strength Equations

Tensile Strength of the Threaded Section 1FT) Tensile strength is determined by multiplying the crosssectional area of the threads by the yield or ultimate strength of the material (depending on whether you define strength as yield or rupture). (See ULTIMATE STRENGTH (Su) or YIELD STRENGTH (Sy) for data.)

Tensile strength (Ff) is usually computed on the hypothetical "tensile stress area" of the threads (As). These areas are tabulated under THREAD STRESS AREAS.

FT ~ SuAs

Shear Strength of the Threaded Section Shear strength is computed by multiplying the tensile stress (As) or root (Ar) cross-sections (see THREAD STRESS AREAS) by the shear strength (Fss) of the material. Shear strengths are approximately 60% of ultimate tensile strengths (see ULTIMATE STRENGTH).

Fss ~ Ss As

Shear Strength of the Body IFsb) Multiply the cross-sectional area of the body (Ab, based On nominal diameter, D) by the shear strength of the material.

2 Fsb ~ Ss (0.7854 0 2)

STRENGTH Of THREADS


~--

~ ~ ~ ~ ... i .. --r-~ ; ~J.f:~ ~ R!f~

e.::: '=!

!§7.~

~

e,:!iI

~ itf! ~ I !!;!;;; ''4

ii..q:3J eJ . ..:.-"!II

Bolting Procedures Reference Manual' 105

STRESS AREA

(See THREAD STRESS AREAS)

STRESS CORROSION CRACKING (SCC)

Definition

Stress corrosion cracking is a condition in which a fastener that is statically loaded well below the material yield strength can suddenly fail. Three essential conditions must be present for SCC to occur.

o A susceptible material o An unfavorable environment (stress and corrosive agent) o An initiating flaw

, Background

SCC failues have occurred in both pressure boundary bolting and in structural supports. The, investigations of these failures show that the essential conditions existed in each case.

Investigation of the structural bolting failures showed that the preloads were high and the bolting material was too hard. The pressure boundary failures were precipitated by jagged thread roots which caused stress concentration, leaks and sealing repairs which provided a corrosive environment.

Basic Techniques for Fighting see Use a low strength material which is generally not susceptible to stress corrosion cracking.

Limit hardness of fastener material to less than 40 HRC, when the environment is no worse than humid air. For more severe environments hardness may be limited even further. For example, A193 B7M materials, used in hydrogen sulfide environments, are limited to a hardness of 94-99 HRB with 100% hardness testing.

Impose supplemental specifications on the fasteners as suggested by the Joint AIF/MPC Task Group on Bolting. This specification includes

10& Bolting l"rocedures Reference Manual

incominginspection, increased sampling for hardness, visual and NDE examinations for defects.

Reduce stress levels. Eliminate stress concentrations due to thread root radius and thread defects (pits, tears).

Housekeeping. Keep corrosive fluids off the joint members by preventing leaks, and cleaning up spills.

Things to Try First Stop leaks as soon as possible; then clean flanges and bolts if possible.

Study preload (torque) specifications. Reduce, if you're not concerned about other problems such as leaks, vibration loosening or fatigue, where high preload is necessary.

Avoid use of molydisulphide LUBRICANTS. The sulphides in these lubricants encourage stress corrosion cracking. .

Avoid use of leak SEALANTS which may (trap electrolytes and release chlorides and sulphides).

Inspect bolts for cracks, using magnetic particle, dye penetrant or ultrasonic techniques.

Need more? Replace bolts. Use materials having lower ultimate strengths.

Use leak detection equipment to spot leaks sooner. Switch to corrosion resistant bolts (type depends on electrolyte

involved-see expert). Last resort: Have a designer determine the threshold stress for SCC,

using a fracture mechanics approach. STRESS CORROSION CRACKING can be avoided by limiting the

preload to sorrie threshold stress level. Recent developments in fracture mechanics have provided a method for calculating this threshold stress. This technique requires some simplifying assumptions. The method provides only approximate solutions, but it has been correlated against field data and has been used to develop preload specifications. See an expert in fracture mechanics.

References The following documents and texts have helped us prepare this section-and can give you additi0!lal information.

1. Chung, Yun. Threshold Preload Levels for AVOiding Stress Corrosion Cracking in High Strength Bolts. Report for Bechtel Power Management. California, January 1984.

2. Merrick, E., A. Rivers, J. Bickford and T. Marston. Prevention of Bolting Degradation and Failure in Pressure Boundary and Support Applications. Paper given at SMIRT-8 Conference, September 1985.

/~'.

Ef·-:; ~

~!1

~!

~." ,

~~

fi!:~~

~

~ ~~

~b~

~ Jii .~ 5i!

~

f!!: ;'~

~i:'3

Jiirl:;'~

Ii!.,;: ~ r-'t t ''iI

~..::;3

~:, -~ -"

pr--!-.-~

~~"'-

---------------------,.c--- 'j--Bolting Procedures Reference Manu._ /107

3. Utility Recommendations and Guidelines for the Purchase Specification and Receipt/Preinstallation Inspection Requirements for ASME Section III, AISC, ANSI/ ASME B31.1, and ANSI B31.5 Bolts and Threaded Fasteners, prepared by the Joint AIF/MPC Task Group on Bolting. May 1985.

STRESS RELAXATION

(See THERMAL EFFECTS)

STRETCH OF FASTENERS

(See ASSEMBLY PROCEDURES-Stretch Control) The change in length or elongation of the fastener as it is tightened.

Stretch can be used to estimate the PRELOAD in the fastener.

Stretch Measurement Validity

Measurement of bolt stretch is an accurate indicator of preload, provided the actual stretch is measured and the measurement is made with sufficient accuracy. Stretch measurement is not a perfect measure of preload, but it is the best method available. The validity of this statement is evidenced by the use of stretch measurements to control preload in many critical assemblies such as nuclear reactor pressure heads.

Stretch measurements are not a direct measure of load. The Elongation Chart gives the approximate stretch for various materials at different stress levels.


TABLE J Elongation Chart for Common Bolting Materials (In Thousandths of an Inch per Inch of Grip Length)

20% 40% 60% 80% 100%

BOLTING of of of of of MATERIAL Yield Yield Yield Yield Yield

ASTM A193 B8, B8M, B8C E = 28.5 X 106 0,2 0,4 0,6 0,8 1.0 30k psi y.s. MONEL 40K psi y.s. 0,3 0,5 0,8 1,1 1,3 SAE GR 2 55K psi y.s. 0,4 0,7 1,1 1,5 1,8 SAE GR 3 80K psi y,s, 0,5 1,1 1,6 2,1 2.7 SAE GR 5, A325 96K psi y,s, 0,6 1,3 1,9 2,6 3,2 ASTM A193 B7, B16 105k psi y.s. 0,7 1.4 2,1 2,8 3,5 SAE GR 8, A490 120K psi y,s, 0,8 1.6 2,4 3,2 4,0 INCONEL 718 180K psi y.s. 1,2 2,4 3,6 4,9 6,1 4340 STEEL, RC47 200K psi y,s, 1,3 2,7 4,0 5,3 6,6 BEST AVAILABLE HIGH STRENGTH BOLT MATERIAL 240K psi y.s. 1.6 3,2 4,8 6,4 8,0 TITANIUM (6AL4V) 134K psi y,s, E = 17 X 106 1.6 3,2 4,8 6,4 8,0

Indicated elongation figures are for various percentages of yield strengths of different bolts per inch of grip length, 6 (Modulus of Elasidty assumed to be 30 x 10unless otherwise noted,) To obtain desired elongation for a particular metal, read the elongation figure in the box under the appropriate percentage of yield and multiply by the grip length, in inches.

For example, to obtain the expected elongation for a SAE Grade 5 bolt 'stretched to 80% of yield, with a 5" grip length, select the appropriate boxed figure, which in this case is 2.6, and multiply by 5. The anSWer is 13 thousandths of an inch.

The Table above gives only an approximation of stretch. See below for a more precise approach.

Bolting Procedures Reference M~;U;:al 109

Calculating Stretch

When using stretch measurements to infer load, we are assuming that the load is uniaxial and that the material behaves elastically according to Hooke's Law:

Txx - E L>.L1L Where: Txx - Tensile stress

E - Modulus of elasticity AL = Tensile strain L - Original unstrained length

Re-writing: Fp/A- E L>.L1L

Where: Fp - Axial load A = Cross-sectional area L>.L - Stretch L - Original unstrained length

The terms of this expression are easily defined, measured, and experimentally verified for a tensile specimen shown in Fig. 27.

FIGURE 27 Tensile Specimen

For a threaded fastener there are varied cross-sectional areas, lengths and stresses to consider, as shown in Fig. 28.

11 0 Bolti~g Procedures Reference Manual

FIGURE 28 Stress Distributions in a Fastener.

, ITr-LS~ILTI g lIIllIllY,' ,

II A. B.

Figure 28 shows the actual stress distribution (A) and an idealized representation (B) of the tensile stress. The effective stressed length (LE) is commonly used to calculate the stretch expected for a given axial load on a fastener.

The various sections of the threaded fastener act like springs in series and each section must be considered.

FIGURE 29 Illustration of a boll having five body sections and one thread section (L6).

§::r1~~n:O LS L6 1

The stretch/load relationship for this type of configuration is: Ll.L = (Fp/E)(LlI A1 + L21 A2 + L31 A3 + ... L61 A6)

For the more common nut and bolt configuration, the stretch/load relationship is:

Ll.L = (Fp/E)[LB/AB + (LG-LB)/As + (HN1+HN2)/(2 As)] We have made a standard assumption here that one-half the nut and

Bolting Procedures Reference Ma~,(.ial 111

one-half the bolt head are in the stressed length of the bolt. This assumption is illustrated in Fig. 20 in LENGTH, EFFECTIVE. The effective length LE is measured from mid-nut to mid-nut.

Stretch Measurement

Micrometers, displacement gages or ultrasonic extensometers may be used to measure stretch.

A "C" type micrometer requires access to both ends of the fastener and a reasonably short fastener length (Fig. 30). Measurements are more accurate if small steel balls are first embedded in the ends (centers) of the bolt.

FIGURE 30 C .. mlcrometer Measurements.

A depth micrometer can be used with a DATUM ROD if the stud is drilled and the datum rod is seated at one end of the stud (see illustration under DATUM ROD). The depth micrometer then shows the stretch of the fastener relative to the datum rod.

Displacement gages, mounted on a reference frame, may be used to measure stretch of the fastener. Both ends of the fastener must be accessible.

An ultrasonic extensometer may be used to measure the stretch of the fastener due to load (Fig. 31).

112 Bolh).~ Procedures Reference Manual

FIGURE 31 Ultrasonic instrument used to measure the bolt stretch.

o+.JfZ}

o ,-,

Each of these methods requires two measurements; one of the unloaded fastener and a measurement after the load is applied. The stretch is calculated from the clifference between the two measurements. Record keeping is therefore imperative.

Practical Considerations for Stretch Control

Accuracy Stretch measurements are the most accurate method of controlling preload on actual joints (see ACCURACY).

Measuring Devices The magnitude of the stretch is usually thousandths of an inch. As a rule-of-thumb, a fastener stretches .001" for each 30,000 psi stress and each 1.0" of effective length. Accurate instruments and techniques must be employed. The resolution of the instrument must be on the order of .0001".

Temperature If the fastener temperature changes between the initial and final measurements, a temperature compensation must be made to the stretch reading.

A 3 deg. F change in temperature of a steel fastener loaded to 50% of yield will produce a 1% difference in the stretch measurements.

-' ~"~~

-------------------- ",--Bolting Procedures Reference Mo.. ..(1 113

When using depth rods, the depth rod should be of the same material as the fasteners and it should be kept at the same temperature as the fastener. This will compensate for the stretch of the fastener which is due to temperature change.

STRIPPED THREADS


STUDS, BROKEN

(See FASTENERS, BROKEN)

TEMPERATURE, HIGH

(See THERMAL EFFECTS)

TENSILE STRENGTH

The maximum tensile load a fastener can support before it breaks. (See ULTIMATE STRENGTH.)

114 Bo._ '6 Procedures Reference Manual

TENSILE STRESS AREA

The hypothetical cross-sectional area of the threaded region of a fastener; used to compute average tensile stress in the fastener, proof loads, etc. It is based upon the mean of the pitch diameter and minor diameter of the threads. (See THREAD STRESS AREAS for Table.)

TENSIONERS

(See ASSEMBLY PROCEDURES-Tensioning; TENSIONINGHydraulic)

A tool, usually hydraulic, used to tighten a bolt by pulling onit rather than torquing it.

TENSIONING, HYDRAULIC

Hydraulic tensioners are widely used to preload large threaded fasteners. They are generally used for fasteners of greater than 11/2" diameter. Many people believe that tensioners provide near-perfect preload control, since the tool's hydraulic ram exerts a controlled and accurate force on the fastener during the assembly operation. Unfortunately, however, the fastener does not retain all of this load when the tensioner is removed. This loss of load is referred to as "tensioner efficiency' I or " elastic recovery' I. A review of the tensioning process is needed to identify the factors which affect the tensioner efficiency.

Tensioning Process

Figure32 is a cross-section of a typical hydraulic tensioner. The main features of the unit are the puller bar, annular hydraulic piston, nut rundown mechanism and tensioner base.

Bolting Procedures Referencl:: ~"'l~nual 115

FIGURE 32 Cut-away view of a typical hydraulic tensioner.

_ FLUID UNDER PRESSURE

Figure 33 illustrates the tensioning process:

o Tensioner Installation -The tensioning base is positioned over the stud and nut. -The puller bar is run down on the exposed stud threads. -The top of the puller engages the drive piston.

o Tensioning Load Applied -Hydraulic pressure is applied to the piston. -Axial force is applied to the stud, causing it to stretch. -The axial force is reacted by the joint as shown. The jOint

is compressed under the tensioner base, under the lower nut and at the joint interface. Note that the upper nut and the jOint immediately beneath the upper nut are not stressed.

o Nut Rundown -While the tensioner load is still applied, the upper nut is

run down against the jOint.

--~ .• --------------116 Bolting Procedures Reference Manual

-The upper nut and the joint surface are now strained slightly, depending on the r;tagnitude of t~e rundown . torque. This small compressIve stress acts m parallel wIth the larger stress beneath the tensioner base.

o Pressure Release -The hydraulic pressure is released and the tensioner

removed. The upper nut and stud are noW carrying full load. There is embedment of material at the thread surfaces and at the nut bearing surface. These embedments cause a loss of load.

FIGURE 33 Tensioning process

TENS10NER ENGAGES STUD

NUT RUN DOWN

Les

STUD PULLED

TENSIONER REMOVED

Bolting P.rocedures Reference Manual 117

Practical Considerations for Tensioning

Insure that the tensioner has enough load capacity. The tensioner load. will have to be 25-30% higher than the preload desired in the stud.

The most important part of the tensioning process is the nut rundown. If the nut is not run down firmly or if it binds during rundown, zero preload can result.

ThIngs That AHect Nut Rundown (And Therefore Accuracy).

D The nut rundown mechanism should be well constructed. Right angle gear arrangements are preferable. You want high, controlled rundown torque.

D Fine stud threads can cause the nut to bind during rundown. Coarse threads are preferable.

D The tensioner base is very important. It must fit squarely on the joint surface to allow the tensioner to pull along the axis of the stud. Check the base for signs of yielding or distortion. A distorted base can cause interference with the nut, preventing nut rundown.

D The studs should be perpendicular to the joint surface. Non-perpendicularity results in stud bending and binding of the nut during ~undown. Shinuning the tensioner can correct for perpendicularity problems.

THERMAL EFFECTS

(See Also THERMAL STRESSES) A Significant change in temperature (100 deg.(F) or more) of a previ

ously tightened jOint can create several different types of problems.

Loss of Strength of the Material

See YIELD STRENGTH for details of the effect of elevated temperatures on the strength of bolting materials. This change in strength must be considered when choosing bolt materials and/or preloads for high temperature service.

118 BoltIng Procedures Reference Manual

Loss of Preload

A change in temperature can result in loss of preload for one or more of the following reasons:

Modulus of Elasticity Effect

Assuming that bolts and joint members experience the same increase in temperature and have the same modulus of elasticity, the change in preload and clamping force on the joint can be estimated from:

Fp2 ~ Fpl EZ/El Where: EZ ~ Modulus of elasticity at final

temperature (psi) El Modulus of elasticity at initial

temperature (psi) Fp2 Final preload (Ibs) Fpl Initial preload (lbs)

Gasket Creep

Current operating temperatures can be high enough to cause creep of elastomeric and other gasket materials. Unfortunately, there's little hard data on the amount of creep one can expect. (See GASKET CREEP for details.)Preload is lost as the gasket "flows .out from under" the bolt load,

Gasket Hysteresis

If the flange material has a higher COEFFICIENT OF EXPANSION than the bolt material or the flange becomes hotter than the bolts, then the clamping force on the jOint (and the stresses in the bolts) will increase (se'e THERMAL STRESS for details), If this is a transient condition (for example, if the bolts eventually reach the flange temperature), then the increase in clamping force disappears, If everything in the joint were fully elastic, the preload and clamping force would merely return to their original room temperature levels. Unfortunately, however, gaskets (including spiral-wound) exhibit a lot of hysteresis-that is, when compressed and then released, they do not return to their original thickness. They take a permanent set-and this means a permanent reduction in preload. The problem can be especially bad if carbon steel bolts are used in an aluminum or stainless steel flange.

.-.,.~--

Bolting P;rocedures Reference Manual 119

One possible result: a thermal cycle (or several of these) can cause a previously tight joint to start leaking.

FIGURE 34 Illustration of the stresses created in the gasket and in the baits of an heat exchanger during initial room temperature bolt-up, Initial pressurization, as the system Is put on-line and heats up, and finally as it starts to cool down. The Chart at the top shows the current temperature In the channel side, shell side and bolts. The chart at the bottom shows stresses in the gasket and (with a different origin) in the bolts. Note that the stresses are created by the initial bolt-up process, are reduced somewhat at pressurization, Increase during the first portion of the heat-up cycle (when the temperature of the bolts significantly lags that of the flange members), are reduced as the bolt temperature continues to rise,and fall off further If the system is now turned off and starts to cool down.

The eventual loss in gasket stress during heat-up is caused by the fact that the gasket will not recover the compression created by differential expansion between bolt and joint members.

8M

600'

, , , ,

400'

206 100-"

0

0

0

1I /

oV

//

BOLT UP

650'~ :50 00"

C AN~EL -" 'f? ,-. , t;;ru 5I~E /'

- -- SHElL S!DE

I I ,- I

I. 1\ 1;' \

0.0. GASKET 1// \\ f:0 fI. 1.0. l?o;ASKET \ l\

:::-- I \1'---. ~i-- , "

I- -- 0/ BO L~Ofo, '1---

PRESS RIZA HEAT UP 1 HEAT UP 2 COOL

120 ,. ___ mg Procedures Reference Manual

Differential Expansion Effect

Differential expansion creates the increase in preload which leads to subsequent gasket set and loss of preload. It's also possible, of course, for expansion· between bolts and joint members to create an immediate loss in preload. If the coefficient of the bolts is greater than that of the flange members and/or the bolts are hotter than the flange, then the bolts" expand away" from the flange and some preload will be lost.

As an example, steel bolts will expand a little farther than a cast iron flange-if the temperature of each is raised by the same amount. The fact that flanges tend to be hotter than bolts <at least if the joints aren't insulated) often prevents a problem in such cases.

Stress Relaxation

A creep-related phenomenon called stress relaxation can also cause a loss of preload at very high temperatures over a period of time. The chart below shows the stress relaxation response of a number of bolting materials.

References . The following documents and texts have helped us prepare this section-and can give you additional information.

1. Hayashi, K. and A. Chang. Development of a Simple Finite Element Model for Elevated Temperature Bolted Flanged Joint. Paper written for research project sponsored by the Subcommittee on Bolted Flanged Connections of the Pressure Vessel Research Committee of the Welding Research Council. New York. To be published.

2. White, P. E., Mountford Corrosion Resistant FastenersHigh and Low Temperature Stud Bolting. GKN Fasteners Corrosion Lab, Reprint No. 32. .

3. Winter, R. Bolting Document. Tennessee Eastman Corporation. 1980.

J Bolting· Procedures Reference Manual

FIGURE 35 Stress relaxation In various bolt materials when they are subjected to high temperatures. Exposure was for 1,000 hours in each test. As an. example, a carbon steel bolt will lose approximately 30% 01 Its Initial preload II exposed to 300 degrees C lor 1,000 hours. It will lose approximately 90% 01 Its Initial preload II exposed to 400 degrees C for the same length of time.

.00

TEMPERATURE ~c

Source: British Standard BS 4882: 1973; Bolting for Flanges and Pressure Containing Purposes.

121

122 BOrtI'iQ~ Procedures Reference Manual

THERMAL STRESSES

(See Also THERMAL EFFECTS) Differential expansion between bolts and flange members can in

crease or decrease preloads, stresses and the clamping force in the joint. Remember that the joint members are trapped between bolt head and nut or between two nuts or nut and tapped hole, etc. If, because of a change in temperature, the bolts expand more (or shrink less) than the flange members, then stresses and forces will decrease. If the bolts expand less, or shrink more, forces and stresses will increase.

A Thermal Stress Worksheet is provided to help you compute the change in stress in your own application.

THREAD STRESS AREAS .

When computing the average stress in the threaded region of a fastener, one must make some' assumption about the cross-sectional area of the threads. A conservative assumption is to base stress calculations on the area defined by the root diameter of the threads (Ar; in2)

Ar ~ 0.7854 (D-1.3IN), Where: 0 ~ The nominal diameter of the bolt (in)

N ~ Number of threads per inch The ASME Code uses root areas for design purposes. In actual fact, the helical threads strengthen the threaded portion

of the fastener somewhat. Experiments show that one can assume an effective cross-section based on the mean of the pitch and root diameters of the threads. For a UN thread, this area, called the Tensile Stress Area (As; in2)(and widely used by fastener manufacturers, designers, users, etc.), can be computed from:

As ~ 0.7854 (0-0.97 43/N)' Where: (Symbols and units same as above)

Table ~, which follows, lists root and tensile stress areas for coarse threads between 1/4-20 and 4-8.


TABLE K Tensile Siress (As) and Thread Rool (Ar) CrossSecllonal Areas

THREAD As (lN2) Ar (IN2)

112-13 0.142 0.126 9116-12 0.182 0.162 518-11 0.226 0.202 314-10 0.334 0.302 718-9 0.462 0.419 1-8 0.606 0.551 1 118-8 0.790 0.728' 1 114-8 0.969 0.890 1 318-8 1.23 1.16 1 112-8 1.49 1.41 1 518-8 1.78 1.68 1 314-8 2.08 1.98 1 718-8 2.41 2.30 2-8 2.50 2.30 2 114-8 3.56 3.42 2 112-8 4.44 4.29 2314-8 5.43 5.26 3-8 6.51 6.32 3 114-8 7.69 7.49 3 112-8 8.96 8.75 3314-8 10.34 10.11 4-8 11.81 11.57

THREAD STRIPPING

The threads will strip when the axial forces on the fastener exceed the shear strength of the male or female threads. The main factors which determine stripping strength are the size of the fastener, the length of engagement of the threads and the strength of the material from which the fastener is made. A number of secondary factors, discussed later, also can affect stripping strength.

MaterIal Strength

The shear strength of most fastener materials is approximately 60% of the ultimate strength. For numerical data, see HARDNESS OF BOlTS or, for a brief summary list of ultimate strengths, see Material Selection under S1RENGTH OF FASTENERS-Static.


Length of Engagement

Increasing the length of engagement between male and female threads increases the cross-sectional area of the material which must be sheared to strip the threads. As a result, heavy hex nuts or deep tapped holes are less easy to strip than regular nuts or shallow holes.

The effect of length is less than you might expect, however. The first few (in-board) threads that engage each other absorb most of the load transmitted from male to female threads. Therefore, adding more lightly loaded threads at the other end of the nut doesn't increase stripping strength as much as we'd like. But it helps up to a point: to a length of 11/2 diameters or so. Beyond that, it's a waste of material.

It's a good idea to design the fastener so that the bolt or stud will break before the threads strip-since a break is easier to detect. As a general rule of thumb you'll have a "break-before-strip" situation if the cross-sectional stripping area of the threads is twice that of the tensile stress area of the male threads (because the tensile strength is roughly twice the shear strength).

See THREAD STRESS AREAS for tensile stress areas. Table L gives thread stripping areas for a length of engagement equal to one nominal diameter of the fastener. Using this information we can see that, for a 11/4-8, Class 2A thread, a length of engagement equal to about 314 D would develop the full strength of the body, so a heavy hex nut (with a height of D) should never strip. All this, of course, assumes correct materials for both parts, correct tolerances on male and female threads, no previous thread damage, etc.

The Influence of Size

The threads on a large fastener are "longer" per turn and have thicker roots than the threads on a small fastener. This means that the perthread area which must be sheared to strip the threads is greater on the larger fastener and that means greater stripping strength.

Doubling the diameter of a fastener will roughly double the stripping strength per inch of thread engagement. Doubling the diameter also

. typically means doubling the length of engagement as well (a nut height or tapped hole depth equal to one nominal diameter is usually used in industry). As a result, stripping strength usually increases by about four times as fastener diameter is doubled. (Doubling the diameter also increases the tensile strength of the body by a factor of 4, so stripping strength keeps pace with tensile strength as the nominal size is changed.)

Bolting Procedures Reference 'tv1allual 125

Stripping Strength Equations

~ale and female (e~ternal and internal) threads can strip along several different cross-sectlOns, depending upon the relative hardness of nut and bolt materials,. ti~htness of fit, etc. The equations for computing s~engt~ can h<; qUIte mvolved and require, as inputs, detailed thread dimensIons ~hich .are not c~nunon1y used. The necessary equations and the thread dimenslOns requrred to solve them can be found in Standard ANSI B1.1 or in any edition of MACHINERY'S HANDBOOK (Industrial Press, Inc., New York).

An estimate of the preload required to strip the threads (Fs) can be cal~ul~ted by multiplying the shear strength of the material (Ss) by the stnppmg area (At) tabulated in Table L. Note that these areas are for e~ternal threads, and for a length af engagement equal to one nominal diameter of the fastener (appropriate for thick or heavy hex nuts). Regular hex nuts typically have a length of 0.875 D). The areas tabulated, furthermare, are for UNC or UNtlueads. You can estimate the stripping strength af other threads by using the follOWing rules-af-thumb:

o The stripping areas of female threads are 1.3 to 1.5 times the areas of the mating male threads (the multiplier is smaller for larger fasteners).

o The stripping areas of fine pitch external threads (UN F) are a~proximately equal to that for Coarser threads up through d~ameters of 1 7/8 inches. As diameters get larger, the finer pitch (now UN) lose ground. By 4" the stripping area of the UNC thread is 5% greater than that of the UN.

o The stripping areas for fine pitch internal threads are 4-5% less than those for coarser threads for all sizes except those in the 1/1 to 1318" range, where coarse and fine areas are essentially equal.

126 Bolti~gProcedures Reference Manual

TABLE L Stripping Areas lor External Threads (In in' per Length 01 Engagement Equal 10 One Nominal Diameter)

STRiPPiNG AREA Thread Ciass 2A Thread Ciass 3A Thread

1/4-20 5/16-18 3/8-16 7/16-14 1/2-13 9/16-12 5/8-11 3/4-10 :!Jl-9

~-,8 1 1/8-8 1 1/4-8 1 3/8-8 1 1/2-8 1 5/8-8 1 3/4-8 1 7/8-8 2-8 2 1/4-8 2 1/2-8 2 3/4-8 3-8 3 1/4-8 3 1/2-8 33/4-8 4-8

Secondary Factors

0.092 0.147 0.216 0.296 0.390 0.502 0.624 0.908 1.25

1.66 2.13 2.65 3.22 3.86 4.55 5.30 6.09 6.96 8.84

10.95 13.28 15.84 18.62 21.63 24.79 28.28

The following factors also affect stripping strength:

0.096 0.157 0.232 0.321 0.427 0.548 0.681 1.01 1.38 1.82 2.329 2.913 3.55 4.26 5.04 5.03 6.81 7.72 9.83

12.18 14.80 17.67 20.8 24.15 27.79 31.64

o A tighter fit between male and female threads increases stripping strength. When fasteners are coated, allowance is left for the coating thickness; this can contribute to lower stripping strength.

o A thin-walled nut will strip more readily than a thickwalled one (more nut dilation).

o Lubricating the fastener before assembly can also increase dilation and reduce stripping strength by a few percent.

Bolting ~rocedures Reference Manual 127

o Threads strip more readily when bolt and nut material are of equal strength. For optimum safety (to guarantee that the bolt will break before the nut threads will strip), use a nut having a specified PROOF LOAD 20% greater than the bolt's ultimate tensile strength.

o Threads will strip more readily if the fastener is torqued during assembly (instead of tensioned) since torquing also enhances dilation.

TORQUE,BREAKAWAY

The maximum torque required to loosen a fastener; to start the reverse motion of the nut.

TORQUE, CONTROL OF

(See ASSEMBLY PROCEDURES-Torquing; TORQUING, SELECTION OF)

TORQUE LOSS


TORQUE PROCEDURES

(See ASSEMBLY PROCEDURES-Torquing)


TORQUE, RE·STARTING

The torque required to start additional forward motion in a previously tightened nut. Used for post-assembly inspection purposes.

TORQUE RELAXATION


TORQUE, SELECTION OF

General

The question most often asked by people who assemble bolted jOints is "What torque should I use on these bolts"? We know from experience that if we use the "wrong torque" something will go wrong: the bolts will be broken during assembly or the joint will leak or it will shake apart or something.

Newcomers to bolting always expect the I/experts!' to have a quick, but correct, answer to the torque question. Sometimes this is possible. But not always. " The behavior of a bolted joint in service depends only indirectly on how much torque was used. The main issue is whether or not the tension created in the bolts by that torque will clamp the joint members together ;"'th a force great enough to resist failure, but small enough so It won t damage the bolts, joint members, gaskets, etc. If the joint has been overdeslgned (and most have been) and/or the service loads and se",;,ice requirements are modest, then a wide range of bolt loads will sabsfy these conditions. Many different torque values will be "good" .

'-'

',---Bolting Procedures Reference h., .~al 129

In-more critical cases, however, it is impossible to specify a "good" torque until we have carefully determined exactly what range of clamping force will resist the service loads without damaging parts. The more demanding the service conditions, the more difficult it is to determine the correct bolt force, or preload.

Another factor complicates the situation still further. Nearly 80 variables can affect the amount of preload achieved in a bolt by a given torque, so even if we've managed to identify the preload which will be best for that joint, we won't necessarily get that preload when we tighten the bolts one by one. Another collection of variables then modifies those initial preloads further as parts relax, as we tighten adjacent bolts, etc. The result is wide variation in preload (see PRELOAD, LOSS OF).

Again, in practice, most joints are overdesigned or "underchallenged", so it doesn't matter if we don't get the preload we selected. As service conditions become more difficult, however, we'll find that we have to take more and more pains to control the uncertainties in the torque-preload relationship.

Detailed torque selection procedures are given in the section on PRELOAD, SELECTION OF. We've chosen to put them there, rather than here, because selection of preload is the more basic issue. Torquing is the most cornmon way to achieve the desired preload, but TENSIONERS or HEATERS can also used. In any event, see PRELOAD, SELECTION OF, for a variety of ways (graded from simple to difficult) to pick a torque.

You'll note in the preload selection discussion that we end up using the same final equation to pick torque (for each preload selection procedure)-

T=KDFp/12 so where's this" difficulty" we were talking about? The inrease in the difficulty (and the accuracy) of our selection. of torque is not visible in this "tmiversal" equation. It is built into the care we take to determine:

Fp the clamping force or preload which will be acceptable for the joint and

K called the NUT FACTOR, an experimentally determined constant which summarizes the many factors influencing the torque-preload relationship.


TORQUE-TURN PROCEDURES

(See ASSEMBLY PROCEDURES-Turn of Nut)

TRAINING BOLTING PERSONNEL

A number of plants have reported that proper training and supervision of bolting crews led to a significant reduction in bolted jOint problems, The training did not involve new types of tools or new procedures-merely instruction in the correct use of existing tools and current procedures, The usefulness of training is confirmed by an NRC study which shows a significant decrease in the incidence of bolting problems as a plant gains experience-Le" as it le.arns how to do the job better.

No formal training programs have been defined as yet for bolting personneL This is surprising, considering the number of bolted joints in a typical plant. One needs oniy compare this situation with the effort that has been made to define the training needs and qualifications for welded joint personnel to see the disparity, Attempts are under way to change this, '

Resource Materials

The Electric Power Research Institute has prepared three video cassette training films for nuclear plant bolting engineers and mechanics, This Reference Manual is another step, again sponsored by EPR!. Both cassettes and manual should make it easier for management to organize and' conduct in-plant training sessions.

Training Program Recommendations A training program should deal with the following issues:

o The basic behavior and "physics" of bolted joints on an elementary leveL (The cassettes provide this,)

o The proper use of the tools, lubricants, gaskets, etc" selected by the plant (the cassettes address this issue, too),

o How to detect and respond to bolted joint problems (a main focus of this manual),

'---Bolting Procedures Reference Mi>.. . .. tl 131

o The importance of "doing it right" (the cassettes and manual),

Reference The following document has helped us prepare this section-and can give you additional informtion,

1, NUREG-1095, "Evaluation of Responses to IE Bulletin 82-02", Nuclear Regulatory Commission,

TURN OF NUT

(See ASSEMBLY PROCEDURES; TURN OF NUT) A formal tightening procedure which consists of snugging all the

bolts in the joint, followed by measured turn of the nut through an additional half turn or the like; widely used in structural steel assembly operations, Properly used turn-of-nut procedures always tighten the fastener past yield,

ULTIMATE STRENGTH

The maximum load that a test specimen or a fastener supports prior to fracture, To express strength as a stress the load is divided by the original tensile cross-section of the fastener or specimen,

The room temperature ultimate strengths for common bolting materials will be found in Tables in STRENGTH OF BOLTING MATERIALS and in HARDNESS OF FASTENERS,

ULTRASONICS

(See EXTENSOMETER)

132 Bolt. l)rocedures Reference Manual

VIBRATION LOOSENING

Vibration can loosen bolts, often causing complete loss of the nut (and bOlt). VariQus explanations have been given for this, but none of them fully explain all of the problems which have been encountered in the field. Most people agree, however, that the following conditions are required before the nut will fully loosen under vibration:

D The vibratory force must have a component at right angles to the axis of the bolts-a force which, when high enough, will cause the joint members to slip past each other. (Vibratory forces parallel to the bolt axis will partially loosen the bolt, causing it to lose perhaps 10-20% of initial preloadbut only transverse forces will fully loosen it.)

f I There mus t be some slip clearance between male and female threads, and between bolts and jOint members, to allow transverse slip to occur.

If these conditions exist, it may take only six severe slip cycles to fully loosen the nut. This means that thermal cycles or flexing of the structure or joint members themselves can loosen it. More commonly, hundreds or thousands of tiny vibration-induced slip motions do the job.

Here are some ways to fight self-loosening. They are graded roughly in order of increasing cost or complexity:

D Increase thread and joint friction forces. If these are high enough, there'll be no transverse slip and therefore no selfloosening. Friction will increase if you: -eliminate thread lubricants -increase preloads (see PRELOAD, SELECTION OF, Level

4 for guidance) -compensate for any RELAXATION of the fasteners (by an

extra pass with the wrench, for example). This increases residual preload, which in turn determines the size of the friction forces.

D Use anaerobic adhesives to "glue" the male and female threads together. These are available for operating temperatures to 450 deg. F or so. At higher temperatures (up to 2000 deg. F) some people use linseed oil. It carbonizes. The nuts must be hammered to break them loose, but there's no damage to the threads.

BaIting Procedures Reference Man~al 133

D Add collars under bolt head and nut(s) and use longer bolts-the longer the better. (A bolt having a length to diameter ratio of 8:1 or more will "never" loosen, some experts say.)

D Use bolts with fine pitch threads instead of coarse (flatter helix angle helps).

o Use special, vibration resistant nuts such as those with a nylon locking collar or interference fit threads. (Avoid the latter if you also have FATIGUE problems or STRESS CORROSION CRACKING with these bolts. The interference can increase stress concentrations.) See Fig. 25 for an illustration of a nylon collar nut.

D Pin or tack-weld joint members to prevent relative slip. Dowel pins or interference-body bolts can be used (Bethlehem makes one for structural use).

D Prevent or reduce the amount of vibration seen by the joint. Some possible ways to do this: -Add stiffeners to joint members -Add mass to change natural frequencies -Use shock absorbers and/or shock mounts to dampen the

vibration D Have a vibration expert study the system to determine the

causes and possible cures for the vibration. D Reshape joint members to prevent relative slip.

FIGURE 36 The lolnt members have been shaped to resist transverse Slip. This can be a very effective way to light vibration loosening.

D Redesign the joint so that the axis of the fasteners is parallel to the direction of vibration rather than perpendicular to it.

134 Bolli rocedures Reference Manual

WASHERS-CRUSH

Crush washers of the type illustrated below are used to control initial preload in bolts and to indicate, on a previously tightened jOin:, wh~ther or not the bolts were properly tightened. They are used prImarily In

structural steel applications.

FIGURE 37 Crush washer used to measure preload. The washer is interposed between the joint member and a conventional washer as shown In the small sketch.

;.ITTI-:

In practice, the crush washer is placed against the joint with its "bumps" upwards. A flat washer is placed on top of the crush washer and the nut follows.

As the nut is tightened, the gap between the washers gets smaller. When a feeler gage can no longer be inserted between the two washers, the bolt has been tightened by the specified amount.

The bumps on the crush washer take a permanent set; the washer cannot be reused. If relaxation occurs in the bolt after tightening, the washer will not show it. But an inspector can indeed tell at any time after initial tightening whether or not the bolt was tightened. And the accuracy cited for the original preloading process is fairly good-+I-IO%is often cited.

A caution. A number of bolting crews have discovered that the bolts are easier to tighten if the bumps on the washers are first pounded flat with a hammer or ground down on a wheel. No crew should be allowed to take advantage of this technological discovery!

Bolting Procedures Reference 1\.1 ... d.ral 135

WASHERS-PLAIN

Hard washers are recommended or specified in a number of bolting standards. They bring several advantages to the bolted joint:

o By spreading the load placed by the bolt or nut on the joint they increase the ratio between jOint and bolt stiffness (see STIFFNESS OF BOLT AND FASTENER). This can help reduce bolt FATIGUE problems.

o Washers make the interface forces between joint members more uniform. This can improve gasket performance.

o Washers can bridge slotted or oversized holes, facilitating assembly of poorly mated parts. They are frequently used for this purpose on structural steel jobs, for example (and are mandatory in many such applications).

o Washers can Significantly reduce the friction between a turning nut and the jOint members. This reduces the size of the bolting tool reqUired, reduces the torques required and often improves the ACCURACY and repeatability of the torquing operations.

o Washers can prevent damage to soft joint surfaces. o Washers reduce the amount of EMBEDMENT between nut,

bolt and joint members, reducing RELAXATION after tightening.

YIELD STRENGTH

The stress level which will produce a small, previously defined amount of pennanent deformation in the fastener. Defonnation of 0.2% is often used to define yield, for example.

The yield strengths of common bolting materials can be found in the chart below.


FIGURE 38 Yield strength of various bolting materials are a function of temperature <an values are approximate).

,eo

'"

'"

~oo 300 700

BOlt TEMPERATURE (of)

/~.,~~.

-------------- \_-

FIGURE 38 Continued

Bolting Procedures Reference Mai .... al 137

A-ASTM A 193 & A320 GRS 88. 88M 8-ASTM A193 8. A320 GR 8BC C-ASTM A193 87; A320 L43; A540 824 CL 5 UP TO 6" OIA E-ASTM A540 821 CL 1 UP TO 4" OIA F-ASTM A540 821 CL 2 UP TO 4" OIA G-ASTM A540 821 CL 3 UP TO 6" OIA H-ASTM A540 821 CL 4 UP TO 6" OIA I-ASTM A540 824 CL 1 UP TO 8" OIA J-ASTM A540 824 CL 2 UP TO 9.5" OIA K-ASTM A540 824V CL 3 UP TO 11" OIA L-ASTM A540 824 CL 4 UP TO 9.5" OIA

Table 1-1.3 of Appendix I, Section Ill. Division 1 ASME Boiler and Pressure Vessel Code, Nuclear Power Plant Components.

Bolting Procedures Reference Ma, 141

GASKET STRESS WORKSHEET

(For Raised Face, Carbon Steel Flanges)

Joint Identlflcatlon ______________ _

Fastener Size _____ ,Materlal _________ _

Data Required (with symbols and units)

10 of Sealing Surface (10; inches __________ _

OD of Sealing Surface (OD; inches, _________ _

Type of Gasket ________________ _

Contained Pressure (P; psi), ____________ _

Number of Bolts in Joint (n) ____________ _

Recommended Seating Stress (SGI; psi)(Ref. A) _____ _

Recommended Residual Stress/Pressure Ratio (Rm)(Ref. A)

References: (A) Gasket Factors

1. Compute, if required, and enter the nominal bolt preload in each bolt at assembly (Fp; Ibs). For example, if preload is specified by torque (T; ft-lbs) then-

Fp ~ 12 T / K D ~ Ibs (See PRELOAD, CALCULATION OF for details, if uncertain. )

2. Compute the total, nominal clamping force on the gasket at assembly (FGA; Ibs).

FGA ~ n Fp ~ ____________ lbs

142 Boltii'_..., ~'rocedures Reference Manual

3. Compute the full surface sealing area of the gasket

(AC, in') AC ~ 0.7854 ( OO'-lD') ~ in'

4. Now use #2 and #3 above to compute the initial stress on the gasket at assembly (SCI; psi)

SCI ~ FCA/AC ~ ____________ psi Is this seating stress greater than the acceptable stress found in CASKET FACTORS and listed above? If so, continue the calculations with the same nominal preload you've used so far. If not, increase that nominal and recycle to step #2. (Note: If the calculated stress exceeds the listed gasket factor by a large amount, consult the gasket manufacturer to see if it's an acceptable stress.)

New Nominal Fp Ibs 5. Compute the pressure load on the joint (LP; Ibs)

LP ~ 0.7854 P (lD') ~ Ibs 6. The pressure load will partially relieve the joint. Compute

the net clamping force on the joint alter the system has been pressurized (FN; lbs)

FN ~ FCA-LP ~ __________ Ibs 7. Compute the residual stress on the gasket (SC2; psi)

SCR ~ FN/AC ~ psi 8. Compute the ratio between this residual stress and the

contained pressure (Rm)

Rm ~ SCRIP - ____ ,....--, ____ ----, __ _ Does Rm exceed the recommended residual stress/contained pressure ratio as per CASKET FACTORS? If so, your nominal preload is accepted. If not, choose a higher nominal preload and recycle to step #2.

New Nominal Fp Ibs 9. Convert the final nominal preload to nominal torque (T;

It-Ibs)

T ~ K 0 Fp / 12 ~ _________ It-Ibs (See NUT FACTORS for K, if required.)


CalculatIons Made By _____________ _

Date_~ __________________ __

~~

~

~

.' !! ~

f! ~

F-I-~

;-r~

1'1 !ill i' I_~

PI ~

~

~

~

~

~

~

tti!f

~

~ F-I::!it --r

r·1-iil --r ~"T"iiiI --y

------------------------------------------~.~------

Bolting Procedures Reference ~ >lal 145

TORQUE COMPUTATION WORKSHEET -LEVEL 3 Joint Identification ____________________________ _

Fastener Size Material _________ _ (Ref: MATERIALS: IDENTIFICATION OF)

Percent of Yield. The first decision you'll have to make, when using the worksheet, is, "'What percentage of yield is correct for this application?" Since this is a Level 3 procedure, the decision is not a critical one, and can be made by considering "common practice' I , Here are some typical values.

PERCENTAGE APPROPRIATE FOR SITUATIONS OF YIELD SUCH AS THESE

.25 Foundation or anchor bolts which see no cyclic or vibration loads; or any unimportant non-gasketed joint exposed to static loads. *

AO Gasketed pressure boundary joints, including those covered by the ASME Code, which have not given trouble and/or are in "routine service" .

(This recommendation is for carbon steel, raised face flanges. For others see FLANGES, UNCOMMON MATERIALS; FLANGES, O-RING, etc.)

.50 "Average" non-gasketed joint, with no great demands upon it; no particular safety or performance requirements; and you have no design information to suggest a higher or lower preload.

* "High preload problems"-STRESS CORROSION CRACKING, THREAD STRIPPING or joint failure, for example-are less common than "low preload problems". However, they may also require preloads of 25-30%. These are Level 4 problems, however, and should not be solved by use of the above table.


.70 A gasketed or non-gasketed joint with which you've had "low preload" problems in the past (leaks, vibration loosening or fatigue). You want a relatively high, but safe, preload; and are going to use simple torquing procedures to tighten the bolts (procedures expected to give torque-preload scatter of +1-30% or so).

.85 Critical gasketed or non-gasketed joints which have been consistent trouble-makers in the past and require maximum safe preload. You're planning to use datum rods, or ultrasonics, or other means to control the tightening process with exceptional accuracy.

1. List the following factors. Use the references cited to look up the necessary data, if required.

Percentage of yield strength to which fasteners are to be tightened (as a decimal) Ref: The Table above.

NOMINAL DIAMETER OF BOLT

YIELD STRENGTH of Bolt Mat'l.

Tensile Stress Area of the Threads (Ref: THREAD STRESS AREAS)

NUT FACTOR (Ref: NUT FACTOR) Conversion Factor (inch-lbs to ft-Ibs)

M~O. ____ _

D ~ _____ inches

Sy ~ ______ psi

As ~ _____ (in')

K~

C ~ 0.0833

2. Multiply all of the above together to compute the torque.

In computer format:

COMPUTED TORQUE M*D*Sy*As*K*C = __________ It-lbs

Bolting Procedures Reference Mal.. 147

3. If using a pevailing torque fastener, add the prevailing torque to the torque calculated above.

Notes:

Calculation Made By, ___________ _

Date' _________________ _

if ~

;r·iiiil ·~--r·

~~\ -----------------------------Bolting Procedures Reference Ma.. ..... al 149

PRELOAD/TORQUESELECTION WORKSHEET -LEVEL 4

.Joint Identiflcatlon _______________ _

Fastener Size _____ Material _________ _

Data Required (with symbols and units)

a) Nominal preload put in bolts at assembly

(Fp; lbs) lbs Alternate to above: Nominal torque used at assembly

(T; ft·lbs) ________ -- ft·lbs; plus type of

lubricant used ________________ _

NUT FACTOR for lubricant (K) ________ -,-Note: The nominal preload (Fp) is the best estimate of the

preload which is currently being applied to the fasteners. This preload may have been obtained from a number of sources such as: torque table (level 2), a torque/preload calculation (level 3), a vendor recommendation or simply from experience. The accuracy of the preload estimate is not impor· tant at this point, since it is just a starting point for the calculations. If the estimate is poor (either high or low), more iteration may be reqUired to arrive at an acceptable preload.

b) Accuracy of preload developed at assembly ref:

ACCURACY(A;%), __ -------~ ______ --~~~ c) Tensile stress area of bolts (As, in2)(Ref:THREAD STRESS

AREAS) in' d) YIELD STRENGTH of bolts at operating temp

(Syt; psi) _______________ psi


Calculations:

1. Compute, if required, and enter nominal preload (Fp, lbs)

Fp ~ 12 T/(K D)~ lbs 2. Use tool ACCURACY to compute max and min assembly

preloads (Fpmin , Fpmax; lbs)

Fpmin ~ Fp (1-A/l00) ~ ________ lbs

Fpmax ~ Fp (1 + AIlOO) ~ lbs 3. Use the THERMAL STRESS worksheet to compute the

stress which will be created by a change in temperature following assembly. If there is no temperature change, enter zero here. (Thermal stress isSTH; psi)

STH ~ psi 4. Compute the change in tension in the bolts, created by

the temperature change (FT; lbs)

FT ~ STH As ~ ____________ lbs Note: FT can be either positive or negative (See

THERMAL STRESS for details.) 5. Combine #2 and #4 above to compute the preload scatter

as a result of the combined effects of tool accuracy and thermal stress. (FATmax, FATmin; lbs)

FATmax ~ Fpmax + FTH ~ lbs

FATmin ~ Fpmin + FTH ~ lbs Note: If FT is negative (relieves the joint) it will be

subtracted from Fpmin and Fpmax. 6. Check the minimum load (FATmin) computed above

against the sealing loads calculated using the Gasket Stress Workskeet: a. Let FATmin ~ Fp in equation 2 of the Gasket Stress

Worksheet; calculate the clamping force FGA. b. Calculate the residual gasket stress (Rm) using steps

6,7,8 (Gasket Worksheet). c. If Rm is acceptable, continue; if not, adjust nominal

preload and return to step 1 of this worksheet (Preload/Torque Selection Worksheet).

5::::!!"~J

r~ ~i:.=JI

E::i~ ~.~ I ' ~r t::~ Cp E::~ c:~ CL::ji(

L ~-~ E> iii !t~ !i

~~

E- iii "

~

~

~

~

~ ......,-

Bolting Procedures Reference lvlQoual 151

d. Calculate the maximum stress on the gasket at operating conditions (SGM).

SGM ~ FN/AG

e. Compare SGM to the maximum gasket stress (see GASKET FACTORS). If acceptable, continue; if not, return to step 1 of this worksheet.

7. Calculate the maximum bolt stress Smax:

Smax ~ FATmax/As ~ __________ psi

8. Compare maximum stress Smax to 2/3 of the bolt yield strength (Syt) at temperature. If Smax < 2/3 Syt continue. If Smax > 2/3 Syp, return to step 1 of'this procedure, enter new preload, and repeat all calculations.

9. When your choice of nominal preload has passed the tests described in steps #6 and #8 above, list it here.

Final Nominal Fp ~ --_________ lbs

10. If desir~~, convert this preload to the nominal torque to be specified for assembly of the joint (T; ft-lbs).

T=KDFp/12

T = ---------_____ ft-lbs

Computed By: ________________ _

Date _______________________ _

t::::;::BI

C;:!I1

r-,:;;;:;iij

~

~

~

=,.,...9

~

t:;:::!l

Bolting Procedures Reference IVlarmal 153

THERMAL STRESS WORKSHEET

"oint Identification ______________ _ Data Required (with units and symbols) Bolt Data.

Nominal diameter (D) in

Threads per inch (TPI) _____________ _

Body length (LBG) in

Length of threads within grip (LSG) in

Height of nut (HN) in

Height of head (or other nut) HH in

TENSILE STRESS AREA of the bolt As in'

Distance across flats, nut or head (Of) in

Grip length (LG) _______________ in

If bolts and joint have the same COEFFICIENT OF EXPANSION and MODULUS OF ELASTICITY:

COEFFICIENT OF EXPANSION (ab) ______ in/in/oF

MODULUS OF ELASTICITY of each (Eb) at the operating

temp _________ (psi)

Operating temperature of the bolts (Tb) _______ OF

Operating temperature of the joint (Tj) ________ OF

If the modulii and/or coefficients differ

COEFFICIENT OF EXPANSION, bolts (ab) ____ in/in/oF

COEFFICIENT OF EXPANSION, joint (aj) ____ in/in/OF


MODULUS of bolts (Eb) at operating temp ______ psi

MODULUS of joint (Ej) at operating temp ______ psi

Common temperature at assy (To) _________ of

Final bolt temp (Tb) _______________ of

Final joint temp (Tj) ______________ of

Initial Calculations: Effective length of the body of the bolt (LB)

LB = LBG + (HH/2) Effective length of the threads (LS)

LS = LSG + (HN/2) Cross-sectional area of the body (AB)

AB = 0.7854 D2 Effective stressed length of bolt (LE)

LE = LB + LS . for a fully threaded stud:

LE = LS + HN/2

Procedure A

When bolts and joint members have the same coefficient of expansion and the same modulus of elasticity: Compute the stresses created in the bolt by the change in temperature (L:.t) from:

STH = "b (LE-LG) (Tj-Tb)

Compute the resulting increase or decrease in preload and clamping force (FT) from:

FTH = (STH)As

Procedure B

When modulii and/or coefficients differ: Compute the expansion (or contraction) of bolts (L:.L) and joint members (L:.J) from:

L:.L = "b LE (Tb-To) = ____________ in

L:.j = "j LG (Tj-To) = ____________ in

Bolting Procedures Reference .,_, .. dlllal 155

Compute the stiffness of the bolts (Kb); Ibs/in -If fully threaded

Kb=EAs LE

-If not fully threaded

Kb = E As AB LS AB + LB As

In calculator format this latter becomes:

(Eb*As*AB)/«LS*AB) + (LB*As)) = Kb

Compute the effective cross-sectional area of that portion of the joint which is loaded by one bolt.

Aj = 2.. [(Df + LG), - D2] 4 2

In calculator format this becomes: 0.7854*«Df + LG/2)2 - D2) = Aj

Compute the stiffness of the jOint (Kj; lbs/in)

Kj = Ej Aj/LG = _____________ Ibs/in

Now compute the change in preload in the bolts (and clamping force on the joint) (Fj; Ibs). If there is a gasket, the gasket stiffness (Kg) is also introduced at this point.

FTH = (L:.L - L:.j) Kb Kj Kb + Kj

In calculator format this becomes:

«L:.L - L:.j)*(Kb*Kj)) I (Kb + Kj) = FTH

Calculations Made By ______________ _

Date ___________________ _

EPRI NP-5067 Good Bolting Practices

Documents

Transcript of EPRI NP-5067 Good Bolting Practices