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SI Prefixes
Multiplication Factor Prefix † Symbol
1 000 000 000 000 5 1012 tera T 1 000 000 000 5 109 giga G
1 000 000 5 106 mega M
1 000 5 103 kilo k
100 5 102 hecto‡ h
10 5 101 deka ‡ da
0.1 5 1021 deci ‡ d
0.01 5 1022 centi ‡ c
0.001 5 1023 milli m
0.000 001 5 1026 micro m
0.000 000 001 5 1029 nano n
212
0.000 000 000 001 5 10 pico p 0.000 000 000 000 001 5 10215
femto f 0.000 000 000 000 000 001 5 10218 atto a
† The first syllable of every prefix is accented so that the prefix will retain its identity.
Thus, the preferred pronunciation of kilometer places the accent on the first syllable,
not the second.
‡ The use of these prefixes should be avoided, except for the measurement of
areas and vol umes and for the nontechnical use of centimeter, as for body and
clothing measurements.
Principal SI Units Used in Mechanics
Quantity Unit Symbol Formula Acceleration
p
m/s2 Angle
squared
p
Meter per second squared
Radian rad † Angular acceleration Radian per second
rad/s2 Angular velocity Radian per second
p
p
rad/s Area
p
Square meter m2 Density Kilogram per cubic meter kg/m3 Energy
Joule J N ? m Force Newton N kg ? m/s2 Frequency Hertz Hz s21
p
Impulse Newton-second kg ? m/s Length Meter m ‡ Mass
p
Kilogram kg ‡ Moment of a force Newton-meter N ? m Power
2
Watt W J/s Pressure Pascal Pa N/m Stress Pascal Pa N/m2 Time
p
Second s ‡ Velocity Meter per second m/s Volume, solids Cubic
p
meter m3 Liquids Liter L 1023 m3 Work Joule J N ? m
† Supplementary unit (1 revolution 5 2p rad 5 3608).
‡ Base unit.
U.S. Customary Units and Their SI Equivalents
Quantity U.S. Customary Units SI Equivalent
Acceleration ft/s2 0.3048 m/s2 in./s2 0.0254 m/s2 Area ft2 0.0929 m2
in2 645.2 mm2 Energy ft ? lb 1.356 J Force kip 4.448 kN lb 4.448
N oz 0.2780 N Impulse lb ? s 4.448 N ? s Length ft 0.3048 m in.
25.40 mm mi 1.609 km Mass oz mass 28.35 g lb mass 0.4536
kg slug 14.59 kg ton 907.2 kg Moment of a force lb ? ft 1.356 N
? m lb ? in. 0.1130 N ? m Moment of inertia
Of an area in4 0.4162 3 106 mm4 Of a mass lb ? ft ? s2 1.356 kg ? m2
Power ft ? lb/s 1.356 W hp 745.7 W Pressure or stress lb/ft2 47.88 Pa
lb/in2 (psi) 6.895 kPa Velocity ft/s 0.3048 m/s in./s 0.0254 m/s mi/h
(mph) 0.4470 m/s
mi/h (mph) 1.609 km/h
Volume, solids ft3 0.02832 m3 in3 16.39 cm3 Liquids gal 3.785 L
qt 0.9464 L Work ft ? lb 1.356 J
MECHANICS OF
MATERIALS
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SIXTH EDITION
MECHANICS OF
MATERIALS
Ferdinand P. Beer
Late of Lehigh University
E. Russell Johnston, Jr.
Late of University of Connecticut
John T. Dewolf
University of Connecticut
David F. Mazurek
United States Coast Guard Academy
TM
TM
MECHANICS OF MATERIALS, SIXTH EDITION
Published by McGraw-Hill, a business unit of The McGraw-Hill Companies, Inc., 1221 Avenue of the Americas, New
York, NY 10020. Copyright © 2012 by The McGraw-Hill Companies, Inc. All rights reserved. Previous editions © 2009,
2006, and 2002. No part of this publication may be reproduced or distributed in any form or by any means, or stored in
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limited to, in any network or other electronic storage or transmission, or broadcast for distance learning.
Some ancillaries, including electronic and print components, may not be available to customers outside the United
States. This book is printed on acid-free paper.
1 2 3 4 5 6 7 8 9 0 QVR/QVR 1 0 9 8 7 6 5 4 3 2 1
ISBN 978-0-07-338028-5
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Library of Congress Cataloging-in-Publication Data
Mechanics of materials / Ferdinand Beer ... [et al.]. — 6th ed.
p. cm.
Includes index.
ISBN 978-0-07-338028-5
ISBN 0-07-338028-8 (alk. paper)
1. Strength of materials—Textbooks. I. Beer, Ferdinand Pierre, 1915–
TA405.B39 2012
620.1’12—dc22
2010037852
www.mhhe.com
About the Authors
As publishers of the books written by Ferd Beer and Russ John
ston, we are often asked how did they happen to write the books
together, with one of them at Lehigh and the other at the University
of Connecticut.
The answer to this question is simple. Russ Johnston’s first teach
ing appointment was in the Department of Civil Engineering and Me
chanics at Lehigh University. There he met Ferd Beer, who had joined
that department two years earlier and was in charge of the courses in
mechanics. Born in France and educated in France and Switzerland
(he held an M.S. degree from the Sorbonne and an Sc.D. degree in the
field of theoretical mechanics from the University of Geneva), Ferd
had come to the United States after serving in the French army during
the early part of World War II and had taught for four years at Williams
College in the Williams-MIT joint arts and engineering program. Born
in Philadelphia, Russ had obtained a B.S. degree in civil engineering
from the University of Delaware and an Sc.D. degree in the field of
structural engineering from MIT.
Ferd was delighted to discover that the young man who had
been hired chiefly to teach graduate structural engineering courses
was not only willing but eager to help him reorganize the mechanics
courses. Both believed that these courses should be taught from a few
basic principles and that the various concepts involved would be best
understood and remembered by the students if they were presented
to them in a graphic way. Together they wrote lecture notes in statics
and dynamics, to which they later added problems they felt would
appeal to future engineers, and soon they produced the manuscript
of the first edition of Mechanics for Engineers. The second edition of
Mechanics for Engineers and the first edition of Vector Mechanics for
Engineers found Russ Johnston at Worcester Polytechnic Institute and
the next editions at the University of Connecticut. In the meantime,
both Ferd and Russ had assumed administrative responsibilities in
their departments, and both were involved in research, consulting,
and supervising graduate students—Ferd in the area of stochastic pro
cesses and random vibrations, and Russ in the area of elastic stability
and structural analysis and design. However, their interest in improv
ing the teaching of the basic mechanics courses had not subsided, and
they both taught sections of these courses as they kept revising their
texts and began writing together the manuscript of the first edition of
Mechanics of Materials.
Ferd and Russ’s contributions to engineering education earned
them a number of honors and awards. They were presented with the
Western Electric Fund Award for excellence in the instruction of en
gineering students by their respective regional sections of the Ameri
can Society for Engineering Education, and they both received the
Distinguished Educator Award from the Mechanics Division of the
v
Johnston team as an author on the second edition of
Mechanics of Materials. John holds a B.S. de gree in
same society. In 1991 Russ received the
civil engineering from the University of Hawaii and
Outstanding Civil Engineer Award from the
Connecticut Section of the American Society of Civil M.E. and Ph.D. degrees in structural engineering
from Cornell University. His research interests are in
Engineers, and in 1995 Ferd was awarded an
honorary Doctor of En gineering degree by Lehigh the area of elastic stability, bridge monitor ing, and
structural analysis and design. He is a registered
University.
Professional Engineering and a member of the
John T. DeWolf, Professor of Civil Engineering at
the University of Connecticut, joined the Beer and Connecticut Board of Professional Engineers. He
vi About the Authors
was selected as the University of Connecticut
Professional Engineer. He has served on the
Teaching Fellow in 2006.
American Railway Engineering & Maintenance of
David F. Mazurek, Professor of Civil Engineering at Way Association’s Commit
the United States Coast Guard Academy, joined the tee 15—Steel Structures for the past seventeen
team in the fourth edition. David holds a B.S. degree years. Professional interests include bridge
in ocean engineering and an M.S. degree in civil
engineering, structural forensics, and blast resistant
engineering from the Florida Institute of Technology, design.
and a Ph.D. degree in civil engineering from the
University of Connecticut. He is a registered
Contents
Preface xii
List of Symbols xviii
1 Introduction—Concept of Stress 2
1.1 Introduction 4
1.2 A Short Review of the Methods of Statics 4
1.3 Stresses in the Members of a Structure 7
1.4 Analysis and Design 8
1.5 Axial Loading; Normal Stress 9
1.6 Shearing Stress 11
1.7 Bearing Stress in Connections 13
1.8 Application to the Analysis and Design of Simple
Structures 13
1.9 Method of Problem Solution 16
1.10 Numerical Accuracy 17
1.11 Stress on an Oblique Plane under Axial Loading 26
1.12 Stress under General Loading Conditions;
Components of Stress 27
1.13 Design Considerations 30
Review and Summary for Chapter 1 42
2 Stress and Strain—Axial
Loading 52
2.1 Introduction 54
2.2 Normal Strain under Axial Loading 55
2.3 Stress-Strain Diagram 57
*2.4 True Stress and True Strain 61
2.5 Hooke’s Law; Modulus of Elasticity 62
2.6 Elastic versus Plastic Behavior of a Material 64
2.7 Repeated Loadings; Fatigue 66
2.8 Deformations of Members under Axial Loading 67
2.9 Statically Indeterminate Problems 78
2.10 Problems Involving Temperature Changes 82
2.11 Poisson’s Ratio 93
2.12 Multiaxial Loading; Generalized Hooke’s Law 94
*2.13 Dilatation; Bulk Modulus 96
vii
viii
Contents 2.14
Shearing Strain 98
2.15 Further Discussion of Deformations under Axial Loading;
Relation among E, n, and G 101
*2.16 Stress-Strain Relationships for Fiber-Reinforced Composite
Materials 103
2.17 Stress and Strain Distribution under Axial Loading;
Saint-Venant’s Principle 113
2.18 Stress Concentrations 115
2.19 Plastic Deformations 117
*2.20 Residual Stresses 121
Review and Summary for Chapter 2 129
3 Torsion 140
3.1 Introduction 142
3.2 Preliminary Discussion of the Stresses in a Shaft 144
3.3 Deformations in a Circular Shaft 145
3.4 Stresses in the Elastic Range 148
3.5 Angle of Twist in the Elastic Range 159
3.6 Statically Indeterminate Shafts 163
3.7 Design of Transmission Shafts 176
3.8 Stress Concentrations in Circular Shafts 179
*3.9 Plastic Deformations in Circular Shafts 184
*3.10 Circular Shafts Made of an Elastoplastic Material 186
*3.11 Residual Stresses in Circular Shafts 189
*3.12 Torsion of Noncircular Members 197
*3.13 Thin-Walled Hollow Shafts 200
Review and Summary for Chapter 3 210
4 Pure Bending 220
4.1 Introduction 222
4.2 Symmetric Member in Pure Bending 224
4.3 Deformations in a Symmetric Member in Pure Bending 226
4.4 Stresses and Deformations in the Elastic Range 229
4.5 Deformations in a Transverse Cross Section 233
4.6 Bending of Members Made of Several Materials 242
4.7 Stress Concentrations 246
*4.8 Plastic Deformations 255
*4.9 Members Made of an Elastoplastic Material 256
*4.10 Plastic Deformations of Members with a Single Plane of
Symmetry 260
*4.11 Residual Stresses 261
ix Contents 4.13 Unsymmetric Bending 279
4.12 Eccentric Axial Loading in a Plane of Symmetry 270
4.14 General Case of Eccentric Axial Loading 284
*4.15 Bending of Curved Members 294
Review and Summary for Chapter 4 305
5 Analysis and Design of Beams for
Bending 314
5.1 Introduction 316
5.2 Shear and Bending-Moment Diagrams 319
5.3 Relations among Load, Shear, and Bending Moment 329
5.4 Design of Prismatic Beams for Bending 339
*5.5 Using Singularity Functions to Determine Shear and Bending
Moment in a Beam 350
*5.6 Nonprismatic Beams 361
Review and Summary for Chapter 5 370
6 Shearing Stresses in Beams and
Thin-Walled Members 380
6.1 Introduction 382
6.2 Shear on the Horizontal Face of a Beam Element 384
6.3 Determination of the Shearing Stresses in a Beam 386
6.4 Shearing Stresses txy in Common Types of Beams 387
*6.5 Further Discussion of the Distribution of Stresses in a
Narrow Rectangular Beam 390
6.6 Longitudinal Shear on a Beam Element of Arbitrary
Shape 399
6.7 Shearing Stresses in Thin-Walled Members 401
*6.8 Plastic Deformations 404
*6.9 Unsymmetric Loading of Thin-Walled Members;
Shear Center 414
Review and Summary for Chapter 6 427
7 Transformations of Stress and
Strain 436
7.1 Introduction 438
7.2 Transformation of Plane Stress 440
7.3 Principal Stresses: Maximum Shearing Stress 443
7.4 Mohr’s Circle for Plane Stress 452
7.5 General State of Stress 462
x Contents 7.6 Application of Mohr’s Circle to the Three-Dimensional Analysis of Stress 464
*7.7 Yield Criteria for Ductile Materials under Plane Stress 467
*7.8 Fracture Criteria for Brittle Materials under Plane Stress 469
7.9 Stresses in Thin-Walled Pressure Vessels 478
*7.10 Transformation of Plane Strain 486
*7.11 Mohr’s Circle for Plane Strain 489
*7.12 Three-Dimensional Analysis of Strain 491
*7.13 Measurements of Strain; Strain Rosette 494
Review and Summary for Chapter 7 502
8 Principal Stresses under a Given
Loading 512
*8.1 Introduction 514
*8.2 Principal Stresses in a Beam 515
*8.3 Design of Transmission Shafts 518
*8.4 Stresses under Combined Loadings 527
Review and Summary for Chapter 8 540
9 Deflection of Beams 548
9.1 Introduction 550
9.2 Deformation of a Beam under Transverse Loading 552
9.3 Equation of the Elastic Curve 553
*9.4 Direct Determination of the Elastic Curve from the Load
Distribution 559
9.5 Statically Indeterminate Beams 561
*9.6 Using Singularity Functions to Determine the Slope and
Deflection of a Beam 571
9.7 Method of Superposition 580
9.8 Application of Superposition to Statically Indeterminate
Beams 582
*9.9 Moment-Area Theorems 592
*9.10 Application to Cantilever Beams and Beams with Symmetric
Loadings 595
*9.11 Bending-Moment Diagrams by Parts 597
*9.12 Application of Moment-Area Theorems to Beams with
Unsymmetric Loadings 605
*9.13 Maximum Deflection 607
*9.14 Use of Moment-Area Theorems with Statically Indeterminate
Beams 609
Review and Summary for Chapter 9 618
xi Contents
10.1 Introduction 632
10.2 Stability of Structures 632
10.3 Euler’s Formula for Pin-Ended Columns 635
10.4 Extension of Euler’s Formula to Columns with Other End
Conditions 638
*10.5 Eccentric Loading; the Secant Formula 649
10.6 Design of Columns under a Centric Load 660
10 Columns 630
10.7 Design of Columns under an Eccentric Load 675
Review and Summary for Chapter 10 684
11 Energy Methods 692
11.1 Introduction 694
11.2 Strain Energy 694
11.3 Strain-Energy Density 696
11.4 Elastic Strain Energy for Normal Stresses 698
11.5 Elastic Strain Energy for Shearing Stresses 701
11.6 Strain Energy for a General State of Stress 704
11.7 Impact Loading 716
11.8 Design for Impact Loads 718
11.9 Work and Energy under a Single Load 719
11.10 Deflection under a Single Load by the
Work-Energy Method 722
*11.11 Work and Energy under Several Loads 732
*11.12 Castigliano’s Theorem 734
*11.13 Deflections by Castigliano’s Theorem 736
*11.14 Statically Indeterminate Structures 740
Review and Summary for Chapter 11 750
Appendices A1
A Moments of Areas A2
B Typical Properties of Selected Materials Used in
Engineering A12
C Properties of Rolled-Steel Shapes A16
D Beam Deflections and Slopes A28
E Fundamentals of Engineering Examination A29
Photo Credits C1
Index I1
Answers to Problems An1
Preface
OBJECTIVES
The main objective of a basic mechanics course should be to develop
in the engineering student the ability to analyze a given problem in
a simple and logical manner and to apply to its solution a few fun
damental and well-understood principles. This text is designed for
the first course in mechanics of materials—or strength of materials—
offered to engineering students in the sophomore or junior year. The
authors hope that it will help instructors achieve this goal in that
particular course in the same way that their other texts may have
helped them in statics and dynamics.
GENERAL APPROACH
In this text the study of the mechanics of materials is based on the
understanding of a few basic concepts and on the use of simplified
models. This approach makes it possible to develop all the necessary
formulas in a rational and logical manner, and to clearly indicate the
conditions under which they can be safely applied to the analysis and
design of actual engineering structures and machine components.
Free-body Diagrams Are Used Extensively. Throughout the
text free-body diagrams are used to determine external or internal
forces. The use of “picture equations” will also help the students
understand the superposition of loadings and the resulting stresses
and deformations.
Design Concepts Are Discussed Throughout the Text When
ever Appropriate. A discussion of the application of the factor
of safety to design can be found in Chap. 1, where the concepts of
both allowable stress design and load and resistance factor design are
presented.
A Careful Balance Between SI and U.S. Customary Units Is
Consistently Maintained. Because it is essential that students be
able to handle effectively both SI metric units and U.S. customary
units, half the examples, sample problems, and problems to be
assigned have been stated in SI units and half in U.S. customary
units. Since a large number of problems are available, instructors can
assign problems using each system of units in whatever proportion
they find most desirable for their class.
Optional Sections Offer Advanced or Specialty Topics. Topics
such as residual stresses, torsion of noncircular and thin-walled mem
bers, bending of curved beams, shearing stresses in non-symmetrical
xii
members, and failure criteria, have been included in optional sec- xiii
Preface
tions for use in courses of varying emphases. To preserve the integ
rity of the subject, these topics are presented in the proper
sequence, wherever they logically belong. Thus, even when not
covered in the course, they are highly visible and can be easily
referred to by the students if needed in a later course or in engi
neering practice. For convenience all optional sections have been
indicated by asterisks.
CHAPTER ORGANIZATION
It is expected that students using this text will have completed a
course in statics. However, Chap. 1 is designed to provide them with
an opportunity to review the concepts learned in that course, while
shear and bending-moment diagrams are covered in detail in Secs.
5.2 and 5.3. The properties of moments and centroids of areas are
described in Appendix A; this material can be used to reinforce the
discussion of the determination of normal and shearing stresses in
beams (Chaps. 4, 5, and 6).
The first four chapters of the text are devoted to the analysis
of the stresses and of the corresponding deformations in various
structural members, considering successively axial loading, torsion,
and pure bending. Each analysis is based on a few basic concepts,
namely, the conditions of equilibrium of the forces exerted on the
member, the relations existing between stress and strain in the mate
rial, and the conditions imposed by the supports and loading of the
member. The study of each type of loading is complemented by a
large number of examples, sample problems, and problems to be
assigned, all designed to strengthen the students’ understanding of
the subject.
The concept of stress at a point is introduced in Chap. 1, where
it is shown that an axial load can produce shearing stresses as well
as normal stresses, depending upon the section considered. The fact
that stresses depend upon the orientation of the surface on which
they are computed is emphasized again in Chaps. 3 and 4 in the
cases of torsion and pure bending. However, the discussion of com
putational techniques—such as Mohr’s circle—used for the transfor
mation of stress at a point is delayed until Chap. 7, after students
have had the opportunity to solve problems involving a combination
of the basic loadings and have discovered for themselves the need
for such techniques.
The discussion in Chap. 2 of the relation between stress and
strain in various materials includes fiber-reinforced composite mate
rials. Also, the study of beams under transverse loads is covered in
two separate chapters. Chapter 5 is devoted to the determination of
the normal stresses in a beam and to the design of beams based
on the allowable normal stress in the material used (Sec. 5.4). The
chapter begins with a discussion of the shear and bending-moment
diagrams (Secs. 5.2 and 5.3) and includes an optional section on the
use of singularity functions for the determination of the shear and
bending moment in a beam (Sec. 5.5). The chapter ends with an
optional section on nonprismatic beams (Sec. 5.6).
method of solution that combines the analysis of
xiv Preface
deformations with the conventional analysis of
Chapter 6 is devoted to the determination of
forces used in statics. In this way, they will have
shearing stresses in beams and thin-walled
members under transverse loadings. The formula for become thoroughly familiar with this fundamental
the shear flow, q5 VQyI, is derived in the traditional method by the end of the course. In addition, this
approach helps the stu
way. More advanced aspects of the design of
dents realize that stresses themselves are statically
beams, such as the determination of the principal
stresses at the junction of the flange and web of a indeterminate and can be computed only by
considering the corresponding distribution of strains.
W-beam, are in Chap. 8, an optional chapter that
The concept of plastic deformation is introduced in
may be covered after the transformations of
Chap. 2, where it is applied to the analysis of
stresses have been discussed in Chap. 7. The
members under axial loading. Problems involving
design of transmission shafts is in that chapter for
the plastic deformation of circular shafts and of
the same reason, as well as the determination of
prismatic beams are also considered in optional
stresses under com bined loadings that can now
sections of Chaps. 3, 4, and 6. While some of this
include the determination of the prin
material can be omitted at the choice of the
cipal stresses, principal planes, and maximum
instructor, its inclusion in the body of the text will help
shearing stress at a given point.
Statically indeterminate problems are first discussed stu
dents realize the limitations of the assumption of a
in Chap. 2 and considered throughout the text for
the various loading conditions encountered. Thus, linear stress-strain relation and serve to caution
them against the inappropriate use of the elastic
students are presented at an early stage with a
torsion and flexure formulas.
PEDAGOGICAL FEATURES
The determination of the deflection of beams is
discussed in Chap. 9. The first part of the chapter is Each chapter begins with an introductory section
devoted to the integration method and to the method setting the purpose and goals of the chapter and
of superposition, with an optional section (Sec. 9.6) describing in simple terms the material to be
covered and its application to the solution of
based on the use of singularity functions. (This
section should be used only if Sec. 5.5 was covered engineering problems.
earlier.) The second part of Chap. 9 is optional. It
presents the moment-area method in two lessons. Chapter Lessons. The body of the text has been
divided into units, each consisting of one or several
Chapter 10 is devoted to columns and contains
material on the design of steel, aluminum, and wood theory sections followed by sample problems and a
large number of problems to be assigned.
columns. Chapter 11 covers energy methods,
including Castigliano’s theorem.
Each unit corresponds to a well-defined topic and generally can be xv
Preface
covered in one lesson.
Examples and Sample Problems. The theory sections include
many examples designed to illustrate the material being presented
and facilitate its understanding. The sample problems are intended
to show some of the applications of the theory to the solution of
engineering problems. Since they have been set up in much the same
form that students will use in solving the assigned problems, the
sample problems serve the double purpose of amplifying the text and
demonstrating the type of neat and orderly work that students should
cultivate in their own solutions.
Homework Problem Sets. Most of the problems are of a practi
cal nature and should appeal to engineering students. They are pri
marily designed, however, to illustrate the material presented in the
text and help the students understand the basic principles used in
mechanics of materials. The problems have been grouped according
to the portions of material they illustrate and have been arranged in
order of increasing difficulty. Problems requiring special attention
have been indicated by asterisks. Answers to problems are given at
the end of the book, except for those with a number set in italics.
Chapter Review and Summary. Each chapter ends with a
review and summary of the material covered in the chapter. Notes
in the margin have been included to help the students organize their
review work, and cross references provided to help them find the
portions of material requiring their special attention.
Review Problems. A set of review problems is included at the end
of each chapter. These problems provide students further opportunity
to apply the most important concepts introduced in the chapter.
Computer Problems. Computers make it possible for engineering
students to solve a great number of challenging problems. A group
of six or more problems designed to be solved with a computer can
be found at the end of each chapter. These problems can be solved
using any computer language that provides a basis for analytical cal
culations. Developing the algorithm required to solve a given problem
will benefit the students in two different ways: (1) it will help them
gain a better understanding of the mechanics principles involved;
(2) it will provide them with an opportunity to apply the skills acquired
in their computer programming course to the solution of a meaning
ful engineering problem. These problems can be solved using any
computer language that provide a basis for analytical calculations.
Fundamentals of Engineering Examination. Engineers who
seek to be licensed as Professional Engineers must take two exams.
The first exam, the Fundamentals of Engineering Examination,
includes subject material from Mechanics of Materials. Appendix E
lists the topics in Mechanics of Materials that are covered in this exam
along with problems that can be solved to review this material.
xvi Preface
SUPPLEMENTAL RESOURCES
Instructor’s Solutions Manual. The Instructor’s and Solutions Manual that accompanies
the sixth edition continues the tradition of exceptional accuracy and keeping solutions
contained to a single page for easier reference. The manual also features tables designed
to assist instructors in creating a schedule of assignments for their courses. The various
topics covered in the text are listed in Table I, and a suggested number of periods to be
spent on each topic is indicated. Table II provides a brief description of all groups of
problems and a classification of the problems in each group according to the units used.
Sample lesson schedules are also found within the manual.
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Preface
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ADDITIONAL ONLINE RESOURCES
Mechanics of Materials 6e also features a companion website (www.
mhhe.com/beerjohnston) for instructors. Included on the website are
lecture PowerPoints, an image library, and animations. Via the website,
instructors can also request access to C.O.S.M.O.S., a complete online
solutions manual organization system that allows instructors to create
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ACKNOWLEDGMENTS
The authors thank the many companies that provided photographs
for this edition. We also wish to recognize the determined efforts
and patience of our photo researcher Sabina Dowell.
Our special thanks go to Professor Dean Updike, of the Depart
ment of Mechanical Engineering and Mechanics, Lehigh University
for his patience and cooperation as he checked the solutions and
answers of all the problems in this edition.
We also gratefully acknowledge the help, comments and sug
gestions offered by the many reviewers and users of previous editions
of Mechanics of Materials.
John T. DeWolf
David F. Mazurek
List of Symbols
a Constant; distance
A, B, C, . . . Forces; reactions
A, B, C, . . . Points
A, A Area
b Distance; width
c Constant; distance; radius
C Centroid
C1, C2, . . . Constants of integration
CP Column stability factor
d Distance; diameter; depth
D Diameter
e Distance; eccentricity; dilatation
E Modulus of elasticity
f Frequency; function
F Force
F.S. Factor of safety
G Modulus of rigidity; shear modulus
h Distance; height
H Force
H, J, K Points
I, Ix, . . . Moment of inertia
Ixy, . . . Product of inertia
J Polar moment of inertia
k Spring constant; shape factor; bulk modulus;
constant
K Stress concentration factor; torsional spring constant
l Length; span
L Length; span
Le Effective length
m Mass
M Couple
M, Mx, . . . Bending moment
MD Bending moment, dead load (LRFD)
ML Bending moment, live load (LRFD)
MU Bending moment, ultimate load (LRFD)
n Number; ratio of moduli of elasticity; normal
direction
p Pressure
P Force; concentrated load
PD Dead load (LRFD)
PL Live load (LRFD)
PU Ultimate load (LRFD)
q Shearing force per unit length; shear flow
Q Force
Q First moment of area
xviii
r Radius; radius of gyration xix
List of Symbols
R Force; reaction
R Radius; modulus of rupture
s Length
S Elastic section modulus
t Thickness; distance; tangential deviation
T Torque
T Temperature
u, v Rectangular coordinates
u Strain-energy density
U Strain energy; work
v Velocity
V Shearing force
V Volume; shear
w Width; distance; load per unit length
W, W Weight, load
x, y, z Rectangular coordinates; distance; displacements;
deflections
x, y, z Coordinates of centroid
Z Plastic section modulus
a, b, g Angles
a Coefficient of thermal expansion; influence
coefficient
g Shearing strain; specific weight
Load factor, dead load (LRFD)
Load factor, live load (LRFD)
gD
gL
displacement
e Normal strain
u Angle; slope
l Direction cosine
n Poisson’s ratio
r Radius of curvature; distance; density
s Normal stress
t Shearing stress
f Angle; angle of twist; resistance factor
v Angular velocity
d Deformation;
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MECHANICS OF
MATERIALS
This chapter is devoted to the study of the stresses occurring in many of the
elements contained in these excavators, such as two-force members, axles,
bolts, and pins.
2
CHAPTER
Introduction—Concept of Stress
3
caused by the application of equal and opposite
transverse forces (Sec. 1.6), and the bearing
1.1 Introduction
stresses created by bolts and pins in the members
1.2 A Short Review of the Methods of Statics
they connect (Sec. 1.7). These various concepts will
1.3 Stresses in the Members of a Structure
be applied in Sec. 1.8 to the determination of the
1.4 Analysis and Design
stresses in the members of the simple structure
1.5 Axial Loading; Normal Stress 1.6 Shearing Stress
considered earlier in Sec. 1.2.
1.7 Bearing Stress in Connections 1.8 Application to the The first part of the chapter ends with a description
Analysis and Design of Simple Structures
of the method you should use in the solution of an
1.9 Method of Problem Solution 1.10 Numerical
assigned problem (Sec. 1.9) and with a discussion
Accuracy
of the numerical accuracy appropriate in
1.11 Stress on an Oblique Plane Under Axial Loading
engineering calculations (Sec. 1.10).
1.12 Stress Under General Loading Conditions;
In Sec. 1.11, where a two-force member under axial
Components of Stress 1.13 Design Considerations
loading is considered again, it will be observed that
1.1 INTRODUCTION
the stresses on an oblique plane include both
The main objective of the study of the mechanics of normal and shearing stresses, while in Sec. 1.12
materials is to provide the future engineer with the you will note that six components are required to
describe the state of stress
means of analyzing and design ing various
at a point in a body under the most general loading
machines and load-bearing structures.
conditions. Finally, Sec. 1.13 will be devoted to the
Both the analysis and the design of a given
structure involve the determination of stresses and determination from test specimens of the ultimate
strength of a given material and to the use of a
deformations. This first chapter is devoted to the
factor of safety in the computation of the allowable
concept of stress.
Section 1.2 is devoted to a short review of the basic load for a structural component made of that
material.
methods of statics and to their application to the
determination of the forces in the members of a
simple structure consisting of pin-connected
1.2 A SHORT REVIEW OF THE METHODS
members. Section 1.3 will introduce you to the
concept of stress in a member of a structure, and OF STATICS
you will be shown how that stress can be
In this section you will review the basic methods of
determined from the force in the member. After a statics while determining the forces in the members
short discussion of engineering analysis and design of a simple structure. Consider the structure shown
(Sec. 1.4), you will consider successively the normal in Fig. 1.1, which was designed to support a 30-kN
stresses in a member under axial loading (Sec. 1.5), load. It consists of a boom AB with a 30 3 50-mm
the shearing stresses
rectangular cross section and of a rod BC with a 20-
Chapter 1 Introduction—Concept of Stress
mm-diameter circular cross section. The boom and Note that the sketch of the structure has been
the rod are connected by a pin at B and are
simplified by omitting all unnec essary details. Many
supported by pins and brackets at A and C,
of you may have recognized at this point that AB
respectively. Our first step should be to draw a free- and BC are two-force members. For those of you
body diagram of the structure by detaching it from who have not, we will pursue our analysis, ignoring
its supports at A and C, and showing the reac tions that fact and assuming that the directions of the
that these supports exert on the structure (Fig. 1.2). reactions at A and C are unknown. Each of these
4
d 20 mm
600 mm
C
50 mm
B
5
1.2 A Short Review of the Methods of Statics
A
800 mm
load.
Fig. 1.1 Boom used to support a 30-kN
30 kN
0
reactions, therefore, will be
represented by two
components, Ax and Ay at A,
and Cx and Cy
at C. We write
the following three equilib
Cy
C
Cx
Ax 5 140 kN (1.1)
0.6 m
1
y o Fx 5 0: Ax 1 Cx 5 0
B
Cx 5 2Ax Cx 5 240 kN (1.2)
rium equations: 1l o MC 5 0:
1x o Fy 5 0: Ay 1 Cy 2 30 kN 5 0 Ay
Ax10.6 m2 2 130 kN210.8 m2 5
obtained from the free- Substituting for Ay from
body diagram of the
(1.4) into (1.3), we
struc ture. We must now obtain Cy 5 130 kN.
Ay 1 Cy 5 130 kN (1.3) dismember the
Expressing the results
We have found two of structure. Considering obtained for the
the four unknowns, but the free body diagram ofreactions at A and C in
cannot determine the the boom AB (Fig. 1.3), vector form, we have
we write the following A A x Fig. 1.2
other two from these
equilibrium equation:
equations, and no
additional independent 1l o MB 5 0: 2Ay10.8 m2
equation can be
5 0 Ay 5 0 (1.4)
0.8 m
Ay By
30 kN
causes tension in that member.
Ax A
B
Bz 0.8 m
30 kN
Fig. 1.3
6 Introduction—Concept of StressFBC FBC
A 5 40 kN y Cx 5 40 kN z, Cy 5 30 kNx
These results could have been anticipated by
We note that the reaction at A is directed along recognizing that AB and BC are two-force
members, i.e., members that are sub jected to
the axis of the boom AB and causes
compression in that member. Observing that the forces at only two points, these points being A
com ponents Cx and Cy of the reaction at C are, and B for member AB, and B and C for member
respectively, proportional to the horizontal and BC. Indeed, for a two-force member the lines of
vertical components of the distance from B to C, action of the resultants of the forces acting at
we conclude that the reaction at C is equal to 50
kN, is directed along the axis of the rod BC, and
each of the two points are obtained a simpler solution
35
equal and opposite and pass by considering the free-body
4
through both points. Using diagram of pin B. The forces
B
30 kN
this property, we could have
Since the force FBC is directed along member
FAB FAB
BC, its slope is the same as that of BC, namely,
3/4. We can, therefore, write the proportion
30 kN
(a) (b)
Fig. 1.4
on pin B are the forces FAB and FBC
exerted,
respectively, by mem
bers AB and BC, and the
30-kN load (Fig. 1.4a). We can express that pin
B is in equilibrium by drawing the corresponding
force triangle (Fig. 1.4b).
4 5 FBC
5 5 30 kN
3
from which we
obtain
FAB
FAB 5 40 kN FBC 5 50 kN
The forces F9AB and F9BC exerted by pin B,
respectively, on boom AB and rod BC are equal
and opposite to FAB and FBC (Fig. 1.5).
FBC
C
FBC
C
D
BC
F'
B BC
F
D
F'BC
Fig. 1.5
B
FAB F'AB A B
Fig. 1.6
F'BC
Knowing the forces at the ends of each of the members, we can
now determine the internal forces in these members. Passing a
section at some arbitrary point D of rod BC, we obtain two por
tions BD and CD (Fig. 1.6). Since 50-kN forces must be applied
at D to both portions of the rod to keep them in equilibrium, we
conclude that an internal force of 50 kN is produced in rod BC
when a 30-kN load is applied at B. We further check from the
directions of the forces FBC and F9BC in Fig. 1.6 that the rod is in
tension. A similar procedure would enable us to determine that
the internal force in boom AB is 40 kN and that the boom is in
compression.
1.3 STRESSES IN THE MEMBERS OF A STRUCTURE 7 1.3 Stresses in the Members of a Structure
While the results obtained in the preceding section represent a first
and necessary step in the analysis of the given structure, they do not
tell us whether the given load can be safely supported. Whether rod
BC, for example, will break or not under this loading depends not
only upon the value found for the internal force FBC, but also upon
the cross-sectional area of the rod and the material of which the rod
will break under the given loading clearly
is made. Indeed, the internal force FBC
depends upon the ability of the material to
actually represents the resul
tant of
withstand the corre sponding value FBCyA of
elementary forces distributed over the entire the intensity of the distributed internal
area A of the cross section (Fig. 1.7) and the
BC BC
average intensity of these distributed A
A
FF
forces is equal to the force per unit area,
FBCyA, in the section. Whether or not the rod
P
P
s 5 A (1.5)
A positive sign will be used to
forces. It thus depends upon the indicate a tensile stress
force FBC, the cross-sectional
P
(member in tension) and a
area A, and the material of the
negative
sign
to
indicate
a
A
rod.
compressive stress (mem ber in
The force per unit area, or
compression).
intensity of the forces distributed Since SI metric units are used
over a given section, is called
in this discussion, with P ex
the stress on that section and is pressed in newtons (N) and A in
denoted by the Greek letter s
square meters (m2), the stress s
(sigma). The stress in a member will be expressed in N/m2. This
of cross-sectional area A
unit is called a pascal (Pa). How
subjected to an axial load P (Fig. ever, one finds that the pascal is
1.8) is therefore obtained by
an exceedingly small quantity P' P'
dividing the magnitude P of the and
load by the area A:
Fig. 1.7 A
that, in practice, multiples of this unit must be
used, namely, the kilopascal (kPa), the
megapascal (MPa), and the gigapascal (GPa).
We have
†The principal SI and U.S. customary units used in mechanics
are listed in tables inside the front cover of this book. From the
table on the right-hand side, we note that 1 psi is approximately
equal to 7 kPa, and 1 ksi approximately equal to 7 MPa.
(a) (b)
Fig. 1.8 Member with an axial load.
1 kPa 5 103 Pa 5 103 N/m2
8 Introduction—Concept of Stress
1.4 ANALYSIS AND DESIGN
1 MPa 5 106 Pa 5 106 N/m2
Considering again the structure of Fig. 1.1, let us
assume that rod BC is made of a steel with a
maximum allowable stress sall 5 165 MPa. Can
When U.S. customary units are used, the force
rod BC safely support the load to which it will be
P is usually expressed in pounds (lb) or
kilopounds (kip), and the cross-sectional area A subjected? The magnitude of the force FBC in
in square inches (in2). The stress s will then be the rod was found earlier to be 50 kN. Recalling
that the diameter of the rod is 20 mm, we use
expressed in pounds per square inch (psi) or
Eq. (1.5) to determine the stress created in the
kilopounds per square inch (ksi).†
rod by the given loading. We have
1 GPa 5 109 Pa 5 109 N/m2
P 5 FBC 5 150 kN 5 150 3 103 N
5 p110 3 1023 m22 5 314 3
A 5 pr2 5 pa20 mm 2 b
1026 m2
2
P
s 5 A 5 150 3 103 N
2
314 3 1026 m 5 1159 3 106 Pa 5 1159 MPa
Since the value obtained for s is smaller than the value sall of the
allowable stress in the steel used, we conclude that rod BC can
safely support the load to which it will be subjected. To be
complete, our analysis of the given structure should also include
the determination of the compressive stress in boom AB, as well
as an investigation of the stresses produced in the pins and their
bearings. This will be discussed later in this chapter. We should
also determine whether the deformations produced by the given
loading are acceptable. The study of deformations under axial
loads will be the subject of Chap. 2. An additional consideration
required for members in compression involves the stability of the
member, i.e., its ability to support a given load without
experiencing a sudden change in configuration. This will be
discussed in Chap. 10.
The engineer’s role is not limited to the analysis of existing
structures and machines subjected to given loading conditions.
Of even greater importance to the engineer is the design of new
struc tures and machines, that is, the selection of appropriate
components to perform a given task. As an example of design,
let us return to the structure of Fig. 1.1, and assume that
aluminum with an allow able stress sall 5 100 MPa is to be used.
Since the force in rod BC will still be P 5 FBC 5 50 kN under the
given loading, we must have, from Eq. (1.5),
sall 5
P
A
P
A 5 sall5 50 3 103 N
100 3 106 Pa 5 500 3 1026 m2
and, since A 5 pr2,
r5 A 5
B p B500 3 1026 m2
p 5 12.62 3 1023 m 5 12.62 mm
d 5 2r 5 25.2 mm
We conclude that an aluminum rod 26 mm or more in diameter
will be adequate.
1.5 AXIAL LOADING; NORMAL STRESS 9 1.5 Axial Loading; Normal Stress
As we have already indicated, rod BC of the example considered in
the preceding section is a two-force member and, therefore, the
forces FBC and F9BC acting on its ends B and C (Fig. 1.5) are directed
along the axis of the rod. We say that the rod is under axial loading.
An actual example of structural members under axial loading is pro
vided by the members of the bridge truss shown in Photo 1.1.
Photo 1.1 This bridge truss consists of two-force members that may be in
tension or in compression.
Returning to rod BC of Fig. 1.5, we recall that the section we
passed through the rod to determine the internal force in the rod
and the corresponding stress was perpendicular to the axis of the
rod; the internal force was therefore normal to the plane of the sec F
tion (Fig. 1.7) and the corresponding stress is described as a normal
stress. Thus, formula (1.5) gives us the normal stress in a
member under axial loading:
A
Q
P
s 5 A (1.5)
We should also note that, in formula (1.5), s is obtained by dividing the
magnitude P of the resultant of the internal forces dis tributed over the
cross section by the area A of the cross section; it represents, therefore,
the average value of the stress over the cross section, rather than the
stress at a specific point of the cross section. P'
To define the stress at a given point Q of the cross section, we
should consider a small area DA
DA approach zero, we obtain the
(Fig. 1.9). Dividing the magnitude of stress at point Q:
DF by DA, we obtain the average
value of the stress over DA. Letting
¢F
Fig. 1.9
¢A (1.6)
s 5 lim ¢Ay0
10 Introduction—Concept of Stress In general, the value obtained for the stress s at a given point Q of the section
is different from the value of the average stress
In a slender rod subjected to equal and
opposite concentrated loads P and P9 (Fig.
1.10a), this variation is small in a section away
from the points of application of the
concentrated loads (Fig. 1.10c), but it is quite
noticeable in the neighborhood of these points
(Fig. 1.10b and d).
given by formula (1.5), and s is found to vary
It follows from Eq. (1.6) that the magnitude of
across the section.
the resultant of the distributed internal forces is
P
#dF 5 #
A
sdA
But the conditions of equilibrium of each of the portions of rod
shown in Fig. 1.10 require that this magnitude be equal to the
mag nitude P of the concentrated loads. We have, therefore,
P5
P' P' P' P' (a) (b) (c) (d)
#dF 5 #
A
s dA (1.7)
ends of the rod. This will be discussed further in
Fig. 1.10 Stress distributions at different sections along axially Chap. 2.
In practice, it will be assumed that the
loaded member.
distribution of normal stresses in an axially
loaded member is uniform, except in the imme
diate vicinity of the points of application of the
loads. The value s of the stress is then equal to
save and can be obtained from formula (1.5).
However, we should realize that, when we
assume a uniform distribution of stresses in the
section, i.e., when we assume that the internal
P
forces are uniformly distributed across the
section, it follows from elementary statics† that
C
the resultant P of the internal forces must be
applied at the centroid C of the section (Fig.
1.11). This means that a uniform distribution of
stress is possible only if the line of action of the
concentrated loads P and P9 passes through the
cen troid of the section considered (Fig. 1.12).
Fig. 1.11
which means that the volume under each of the This type of loading is called centric loading and
stress surfaces in Fig. 1.10 must be equal to the will be assumed to take place in all straight twomagnitude P of the loads. This, how ever, is the force members found in trusses and pinconnected structures, such as the one
only information that we can derive from our
knowledge of statics, regarding the distribution of considered in Fig. 1.1. However, if a two-force
mem ber is loaded axially, but eccentrically as
normal stresses in the various sections of the
shown in Fig. 1.13a, we find from the conditions
rod. The actual distribution of stresses in any
given section is statically indeterminate. To learn of equilibrium of the portion of member shown in
Fig. 1.13b that the internal forces in a given
more about this distribu tion, it is necessary to
section must be
consider the deformations resulting from the
particular mode of application of the loads at the
†See Ferdinand P. Beer and E. Russell Johnston, Jr., Mechanics Mechanics for Engineers, 9th ed., McGraw-Hill, New York, 2010,
for Engineers, 5th ed., McGraw-Hill, New York, 2008, or Vector Secs. 5.2 and 5.3.
P
1.6 Shearing Stress
P
11
C
d
P
CMd
P'
Fig. 1.12
type of stress is obtained when transverse forces
P and P9 are applied to a member AB (Fig.
1.14). Passing a section at C between the points
equivalent to a force P applied at the centroid of of application of the two forces (Fig. 1.15a), we
the section and a couple M of moment M 5 Pd. obtain the diagram of portion AC shown in Fig.
The distribution of forces—and, thus, the
1.15b. We conclude that inter
corresponding distribution of stresses—cannot nal forces must exist in the plane of the section,
be uniform. Nor can the distribution of stresses and that their resul tant is equal to P. These
be symmetric as shown in Fig. 1.10. This point elementary internal forces are called shearing
will be discussed in detail in Chap. 4.
forces, and the magnitude P of their resultant is
the shear in the section. Dividing the shear P by
the area A of the cross section, we
1.6 SHEARING STRESS
P' P'
The internal forces and the corresponding
(a) (b)
stresses discussed in Secs. 1.2 and 1.3 were
normal to the section considered. A very different Fig. 1.13 Eccentric axial loading. P
P
Fig. 1.15
B
P
(a)
AB
AC
P'
Fig. 1.14 Member with
transverse loads.
AC
P'
P
(b)
12 Introduction—Concept of Stress obtain the average shearing stress in the section. Denoting the shear ing
stress by the Greek letter t (tau), we write
tave 5
P
A (1.8)
It should be emphasized that the value obtained is an average
value of the shearing stress over the entire section. Contrary to what
we said earlier for normal stresses, the distribution of shearing
stresses across the section cannot be assumed uniform. As you will
see in Chap. 6, the actual value t of the shearing stress varies from
zero at the surface of the member to a maximum value tmax that may
be much larger than the average value tave.
Photo 1.2 Cutaway view of a connection with a bolt in shear.
Shearing stresses are commonly found in bolts, pins, and rivets
used to connect various structural members and machine compo
nents (Photo 1.2). Consider the two plates A and B, which are con
nected by a bolt CD (Fig. 1.16). If the plates are subjected to tension
forces of magnitude F, stresses will develop in the section of bolt
corresponding to the plane EE9. Drawing the diagrams of the bolt
and of the portion located above the plane EE9 (Fig. 1.17), we con
clude that the shear P in the section is equal to F. The average
shearing stress in the section is obtained, according to formula (1.8),
by dividing the shear P 5 F by the area A of the cross section:
tave 5
CC
F
C
E
F'
F
AF
B E' D
P
F
A 5 A (1.9)
E
E
D
P F'
Fig. 1.16 Bolt subject to single shear.
(a) (b)
Fig. 1.17
13
1.8 Application to the Analysis and
Design of Simple Structures
K
EHC
K'
F' F B A
L
D
L'
GJ
Fig. 1.18 Bolts subject to double shear.
H
The bolt we have just considered is said to be in single shear.
FC
Different loading situations may arise, however. For example, if
P
K
K'
splice plates C and D are used to connect plates A and B (Fig. 1.18),
shear will take place in bolt HJ in average shear ing stress is
bolt CD that we have discussed
each of the two planes KK9 and
in the preceding section (Fig.
LL9 (and similarly in bolt EG). The tave 5 PA 5 Fy2
1.16). The bolt exerts
bolts are said to be in double
shear. To determine the average
Fig. 1.19
shearing stress in each plane, we
(a) (b)
L' F
L
FD
J
F
P
F
A 5 2A (1.10)
1.7 BEARING STRESS IN
CONNECTIONS
draw free-body diagrams of bolt Bolts, pins, and rivets create
stresses in the members they
HJ and of the portion of bolt
located between the two planes connect, along the bearing
surface, or surface of contact. For t
(Fig. 1.19). Observing that the
shear P in each of the sections is example, con sider again the two C
P
plates A and B connected by a
P 5 Fy2, we conclude that the
plate thickness and
on plate A a force P diameter d and of stress, called the
d the diameter of
equal and opposite length t equal to the bearing stress,
thickness
of
the
obtained
by
dividing
the bolt, we have
to the force F
A
plate.
Since
the
the
load
P
by
the
exerted by the plate
distribution of these area of the
on the bolt (Fig.
1.20). The force P forces—and of the rectangle
correspond ing
representing the
represents the
stresses—is quite projection of the bolt
resultant of
on the plate section Fig. 1.20
elementary forces complicated, one
d
distributed on the uses in practice an (Fig. 1.21). Since
Ft
inside surface of a average nominal this area is equal to
td, where t is the
value sb of the
half- cylinder of
F'
D
1.8 APPLICATION TO THE
ANALYSIS AND DESIGN OF
SIMPLE STRUCTURES
We are now in a position to determine the
stresses in the members and
connections of various simple twodimensional structures and, thus, to
design such structures.
A d Fig. 1.21
P
P
sb 5 A 5 td (1.11)
14 Introduction—Concept of Stress As an example, let us return to the structure of Fig. 1.1 that we have already
considered in Sec. 1.2 and let us specify the supports
and connections at A, B, and C. As shown in Fig. 1.22, the 20-mm
diameter rod BC has flat ends of 20 3 40-mm rectangular cross
section, while boom AB has a 30 3 50-mm rectangular cross section
and is fitted with a clevis at end B. Both members are connected at
B by a pin from which the 30-kN load is suspended by means of a
U-shaped bracket. Boom AB is supported at A by a pin fitted into a
double bracket, while rod BC is connected at C to a single bracket.
All pins are 25 mm in diameter.
C
d 25 mm
20 mm
TOP VIEW OF ROD BC Flat end
d 20 mm
Flat end
C
d 25 mm d 20 mm
600 mm
40 mm
B
FRONT VIEW 50 mm
AB B
800 mm
30 mm
Q 30 kN Q 30 kN
END VIEW
20 mm
25 mm
20 mm
25 mm
A
d 25 mm
Fig. 1.22
B
TOP VIEW OF BOOM AB
a. Determination of the Normal Stress in Boom AB and Rod
BC. As we found in Secs. 1.2 and 1.4, the force in rod BC is FBC
5 50 kN (tension) and the area of its circular cross section is A 5
314 3 1026 m2; the corresponding average normal stress is sBC 5
1159 MPa. However, the flat parts of the rod are also under
tension and at the narrowest section, where a hole is located, we
have
A 5 120 mm2140 mm 2 25 mm2 5 300 3 1026 m2
The corresponding average value of the stress, therefore, is 15
1.8 Application to the Analysis and
Design of Simple Structures
1sBC2end 5
P
A 5 50 3 103 N
2
300 3 1026 m 5 167 MPa
Note that this is an average value; close to the hole, the stress will
actually reach a much larger value, as you will see in Sec. 2.18. It is
clear that, under an increasing load, the rod will fail near one of the
holes rather than in rectangular cross
sections of minimum C
section
is
A
5
30
mm
its cylindrical portion;
area at A and B are
its design, therefore, 3 50 mm 5 1.5 3
not under stress,
could be improved 1023 m2, the average since the boom is in
value of the normal compression, and,
by increasing the
stress in the main therefore, pushes on
width or the
(a)
thickness of the flat part of the rod,
between
pins
A
and
ends of the rod.
B, is
Turning now our
attention to boom
AB, we recall from sAB 5 2 40 3 103 N
Sec. 1.2 that the
force in the boom is 1.5 3 1023 m2 5
FAB 5 40 kN
(b)
226.7 3 106 Pa 5
(compression).
D'
Since the area of 226.7 MPa
D
the boom’s
Note that the
50 kN
50 kN
d 25 mm
P
50 kN
Fb
(c)
the pins (instead of pulling on the pins as rod occur (Fig. 1.23c), we conclude that the shear
in that plane is P 5 50 kN. Since the crossBC does).
sectional area of the pin is
b. Determination of the Shearing Stress in Fig. 1.23
Various Connec tions. To determine the
shearing stress in a connection such as a
bolt, pin, or rivet, we first clearly show the
forces exerted by the various members it
connects. Thus, in the case of pin C of our
example (Fig. 1.23a), we draw Fig. 1.23b,
showing the 50-kN force exerted by member
BC on the pin, and the equal and opposite
force exerted by the bracket. Drawing now the A
diagram of the portion of the pin located
below the plane DD9 where shearing stresses
40 kN
A 5 pr2 5 pa25 mm 2 b
5 p112.5 3 1023 m22 5 491 3
1026 m2
2
we find that the average value of the shearing stress in the pin at C is
tave 5
P
A 5 50 3 103 N
2
491 3 1026 m 5 102 MPa
Fb
Considering now the pin at A (Fig. 1.24), we note that it is in
D
double shear. Drawing the free-body diagrams of the pin and of the E
portion of pin located between the planes DD9 and EE9 where shear
Fb
ing stresses occur, we conclude that P 5 20 kN and that tave 5
P
A 5 20 kN
(a)
d 25 mm (b)
D' E'
40 kN
P
P
(c)
40 kN
MPa
Fig. 1.24
2
491 3 1026 m 5 40.7
16 Introduction—Concept of Stress Considering the pin at B (Fig. 1.25a), we note that the pin may be divided into
five portions which are acted upon by forces
exerted by the boom, rod, and bracket. Considering successively
the portions DE (Fig. 1.25b) and DG (Fig. 1.25c), we conclude that
the shear in section E is PE 5 15 kN, while the shear in section G
of the pin is symmetric, we con
1
2 FAB
20 kN
clude that the maximum value of
the shear in pin B is PG 5 25 kN,
where
and that the largest shearing
2 FAB 20 kN
H
stresses occur in sections G and 1
5 25 kN. Since the loading H,
is PG
J
1
Pin B
D
1
E
G
2 Q 15 kN
FBC 50 kN
tave 5 PG
A 5 25 kN
2 50.9 MPa
491 3 1026 m 5
2Q
15 kN
P
td 5 40 kN
D
sb 5
1
130 mm2125 mm25 53.3 MPa
(a)
c. Determination of the Bearing Stresses. To
determine the nominal bearing stress at A in
member AB, we use formula (1.11)
2Q
PE
1
15 kN
(b)
of Sec. 1.7. From Fig. 1.22, we have t 5 30 mm To obtain the bearing stress in the bracket at A,
and d 5 25 mm. Recalling that P 5 FAB 5 40 kN, we use t 5 2(25 mm) 5 50 mm and d 5 25 mm:
we have
E
sb 5
2 FAB
P
td 5 40 kN
member AB, at B and C in mem
20 kN 1
150 mm2125 mm25 32.0 MPa ber BC, and in the bracket at C
are found in a similar way.
PG
G
D
The bearing stresses at B in
“intuition.” After an answer has been obtained, it
2 Q 15 kN
should be checked. Here again, you may call upon
(c)
your common sense and personal experience. If
Fig. 1.25
not completely satisfied with the result obtained,
you should carefully check your formulation of the
1.9 METHOD OF PROBLEM SOLUTION
problem, the validity of the methods used in its
You should approach a problem in mechanics of
solution, and the accuracy of your computations.
materials as you would approach an actual
The statement of the problem should be clear and
engineering situation. By drawing on your own
precise. It should contain the given data and
experience and intuition, you will find it easier to
indicate what information is required. A simplified
understand and formulate the problem. Once the drawing showing all essential quantities involved
problem has been clearly stated, however, there is should be included. The solution of most of the
no place in its solution for your particular fancy.
problems you will encounter will necessitate that
Your solution must be based on the fundamental you first determine the reac
principles of statics and on the principles you will
tions at supports and internal forces and couples.
learn in this course. Every step you take must be
This will require
justified on that basis, leaving no room for your
the drawing of one or several free-body diagrams, as was done in 17
1.10 Numerical Accuracy
Sec. 1.2, from which you will write equilibrium equations. These
equations can be solved for the unknown forces, from which the
required stresses and deformations will be computed.
After the answer has been obtained, it should be carefully
checked. Mistakes in reasoning can often be detected by carrying
the units through your computations and checking the units
obtained for the answer. For example, in the design of the rod
discussed in Sec. 1.4, we found, after carrying the units through
our computations, that the required diameter of the rod was
expressed in millimeters, which is the correct unit for a dimension;
if another unit had been found, we would have known that some
mistake had been made.
Errors in computation will usually be found by substituting the
numerical values obtained into an equation which has not yet been
used and verifying that the equation is satisfied. The importance of
correct computations in engineering cannot be overemphasized.
1.10 NUMERICAL ACCURACY
The accuracy of the solution of a problem depends upon two items:
(1) the accuracy of the given data and (2) the accuracy of the com
putations performed.
The solution cannot be more accurate than the less accurate of
these two items. For example, if the loading of a beam is known to
be 75,000 lb with a possible error of 100 lb either way, the relative
error which measures the degree of accuracy of the data is
100 lb
75,000 lb 5 0.0013 5 0.13%
In computing the reaction at one of the beam supports, it would
then be meaningless to record it as 14,322 lb. The accuracy of the
solution cannot be greater than 0.13%, no matter how accurate the
computations are, and the possible error in the answer may be as
large as (0.13y100)(14,322 lb) < 20 lb. The answer should be prop
erly recorded as 14,320 6 20 lb.
In engineering problems, the data are seldom known with an
accuracy greater than 0.2%. It is therefore seldom justified to write
the answers to such problems with an accuracy greater than 0.2%.
A practical rule is to use 4 figures to record numbers beginning with
a “1” and 3 figures in all other cases. Unless otherwise indicated, the
data given in a problem should be assumed known with a comparable
degree of accuracy. A force of 40 lb, for example, should be read
40.0 lb, and a force of 15 lb should be read 15.00 lb.
Pocket calculators and computers are widely used by practicing
engineers and engineering students. The speed and accuracy of these
devices facilitate the numerical computations in the solution of many
problems. However, students should not record more significant fig
ures than can be justified merely because they are easily obtained.
As noted above, an accuracy greater than 0.2% is seldom necessary
or meaningful in the solution of practical engineering problems.
A
B
D
1.25 in.
SAMPLE PROBLEM
In the hanger shown, the
upper portion of link ABC is 38
in. thick and the lower
portions are each 14 in. thick.
Epoxy resin is used to bond
the upper and lower portions
together at B. The pin at A is
of 38-in. diameter while a 1
4-in.-diameter pin is used at C.
Determine (a) the shearing
stress in pin A, (b) the
shearing stress in pin C, (c)
the largest normal stress in
link ABC, (d) the average
shearing stress on the bonded
surfaces at B, (e) the bearing
1.1
stress in the link at C. 6 in. 1.75 in.
C
7 in.
E
SOLUTION
10 in.
500 lb
Free Body: Entire
Hanger. Since the link
ABC is a two-force
5 in.
AD
Dx
5 in.
10 in.
C
500 lb
1l oMD 5 0: 1500 lb2115 in.2 2
E
FAC Dy
member, the reaction at
A is vertical; the reaction
at D is represented by its
compo nents Dx and Dy.
We write
FAC110 in.2 5 0
FAC 5 1750 lb
lb
FAC 5 750
tension
b. Shearing Stress in
Pin C. Since this 14in.-diameter pin is in
double
a. Shearing Stress in Pin A. Since 750 lb
750 lb
FAC
FAC
this 38-in.-diameter pin is in single
750
lb
A
C
shear, we write
shear, we write
tA 5 FAC
tC 5
A 5 750 lb
1
2FAC
A 5 375 lb
4p10.25
4p10.375
2
in.2 tA 5 6790 psi ◀
2
in.2 tC 5
7640 psi ◀
1
1
-in. diameter 38FAC 375 lb 12FAC 375 lb 12
c. Largest Normal Stress in Link ABC. The largest
14
-in. diameter
where the area is smallest; this
occurs at the cross section at A
where the 38-in. hole is located.
We have
3
in.
8
1.25 in.
stress
is found
0.328
in2 sA 5
2290
psi ◀
sA 5
FAC
Anet5
750 lb
138
Average Shearing Stress at B.
in.211.25 in. 2 0.375 in.25 750 We note that bonding exists on
lb
FAC 750 lb 1.25 in.
-in. diameter 38
FAC
A
F1 F2 FAC 375 lb 12
F2 F1
d.
the shear force on each
side is F1 5 (750 lb)y2 5
375 lb. The average
1.75 in.
shearing stress on each
both sides of the upper
portion of the link and that surface is thus
B
tB 5 F1
11.25 in.211.75 in.2 tB 5 171.4 psi ◀
A 5 375 lb
375 lb
F1 375 lb
1
4 in.
-in. diameter 14
e. Bearing Stress in Link at C. For each portion of
the link, F1 5 375 lb and the nominal bearing area
is (0.25 in.)(0.25 in.) 5 0.0625 in2.
sb 5 F1
A 5 375 lb
2
0.0625 in sb 5 6000 psi ◀
18
AB
F1
F1
d
1
F1 P
SAMPLE PROBLEM 1.2
The steel tie bar shown is to be designed
to carry a tension force of magnitude P 5
120 kN when bolted between double
brackets at A and B. The bar will
be fabricated from 20-mm-thick
plate stock. For the grade of steel
to be used, the maximum
allowable stresses are: s 5 175
MPa, t 5 100 MPa, sb 5
350 MPa. Design the tie bar by
determining the required values of
(a) the diameter d of the bolt, (b)
the dimension b at each end of the
bar, (c) the dimension h of the bar.
a. Diameter of the Bolt. Since the bolt is in
double shear, F1 5 12 P 5 60 kN.
P
4p
SOLUTION
d
2
We will use d 5 28 mm ◀
2
A 5 60 kN
t 20 mm
1
t 5 F1
4p
d
h
d
b
t
1
P
2
a
At this point we
check the bearing
stress between the
20-mm-thick plate
and the 28-mm-
2
100 MPa 5 60 kN
d 5 27.6 mm 1
diameter bolt.
10.020 m210.028 m25 214 MPa , 350 MPa
P
tb 5 td 5 120 kN
OK
b. Dimension b at Each End of the Bar. We consider
one of the end portions of the bar. Recalling that the
thickness of the steel plate is t 5 20 mm and that the
average tensile stress must not exceed 175 MPa, we
write
175 MPa 5 60 kN
ta
d
ba
P' 120 kN P12
17.14 mm
2P
s5
10.02 m2a
a5
1
P 120 kN
b 5 d 1 2a 5 28 mm 1 2(17.14 mm)
b 5 62.3 mm ◀
t 20 mm
h
c. Dimension h of the Bar.
Recalling that the thickness of the
steel plate is t 5 20 mm, we have
10.020 m2h
P
s 5 th
h 5 34.3 mm
We will use h 5 35 mm ◀
175 MPa 5 120 kN
19
PROBLEMS
150 MPa in rod BC,
determine the smallest
allowable values of d1
and d2.
1.2 Two solid cylindrical
rods AB and BC are
welded together at B
2 in.3 in.
and loaded as shown. 30 kN
Knowing that d1 5 50
Fig. P1.1 and P1.2
mm and d2 5 30 mm,
find the average normal
stress at the midsection
of (a) rod AB, (b) rod
P
BC.
C
d1
300 mm 250 mm
40 kN d2
A
P for which the tensile
stress in rod AB has the
same magnitude as the
compressive stress in
rod BC.
1.1 Two solid
C
cylindrical rods AB and
BC are welded together 1.3 Two solid cylindrical A
30 kips
rods AB and BC are
at B and loaded as
B
welded
together
at
B
shown. Knowing that
and
loaded
as
shown.
the average normal
stress must not exceed Determine the
175 MPa in rod AB and magnitude of the force
B
30 kips
30 in. 40 in.
Fig. P1.3
1.4 In Prob. 1.3, knowing that P 5 40 kips, determine the average
normal stress at the midsection of (a) rod AB, (b) rod BC.
1.5 Two steel plates are to be held together by means of 16-mm-diameter
high-strength steel bolts fitting snugly inside cylindrical brass
spacers. Knowing that the average normal stress must not exceed
200 MPa in the bolts and 130 MPa in the spacers, determine the
outer diameter of the spacers that yields the most economical and
safe design.
A
a
15 mm
B
Fig. P1.5
100 m
1.6 Two brass rods AB and BC, each of uniform diameter, will be
b
Fig. P1.6
10 mm
C
brazed together at B to form a
nonuniform rod of total length 100 normal stress in ABC is
m which will be suspended from a minimum, (b) the corresponding
support at A as shown. Knowing value of the maximum normal
that the density of brass is 8470 stress.
kg/m3, determine (a) the length of
rod AB for which the maximum
20
21 Problems 1.7 Each of the four vertical links has an 8 3 36-mm uniform rectan gular cross section and each of the
four pins has a 16-mm diameter.
Determine the maximum value of the average normal stress in the
links connecting (a) points B and D, (b) points C and E.
0.4 m
C
0.25 m
D
B
20 kN
A
0.2 m
4 in.
12 in. 4 in.
E
E
stress in the central portion of that
link when (a) u 5 0,
B
2 in.
C D J 6 in.
Fig. P1.7
60 lb
F
A
8 in.
1.8 Knowing that link DE is 18 in. thick(b) u 5 908.
and 1 in. wide, determine the normal D
1.9 Link AC has a uniform rectangular Fig. P1.8
cross section 116 in. thick and 1
4 in. wide. Determine the normal stress in
the central portion of the link.
Fig. P1.9
3 in.
6 in.
240 lb
7 in.
240 lb
C
BA
30
1.10 Three forces, each of magnitude P 5 4 kN, are applied to the
mechanism shown. Determine the cross-sectional area of the
uni form portion of rod BE for which the normal stress in that
portion is 1100 MPa.
0.100 m
E
PPP
D
ABC
0.150 m 0.300 m 0.250 m
Fig. P1.10
22 Introduction—Concept of Stress 1.11 The frame shown consists of four wooden members, ABC, DEF,
BE, and CF.
Knowing that each member has a 2 3 4-in. rectan
gular cross section and that each pin has a
1 -in.
2
diameter, deter
mine the maximum value of the average normal stress
(a) in
member BE, (b) in member CF.
30 in.
45 in.
AB
40 in.
4 in.
C 480 lb
4 in.
DEF
Fig. P1.11
15 in.
30 in.
1.12 For the Pratt bridge truss and loading shown, determine the
aver age normal stress in member BE, knowing that the crosssectional area of that member is 5.87 in2.
BDF
12 ft
CEG
A
9 ft 9 ft 9 ft
9 ft
H
Fig. P1.12
80 kips 80 kips 80 kips
1.13 An aircraft tow bar is positioned by means of a single hydraulic
cylinder connected by a 25-mm-diameter steel rod to two
identical arm-and-wheel units DEF. The mass of the entire tow
bar is 200 kg, and its center of gravity is located at G. For the
position shown, determine the normal stress in the rod.
1150
A
D
850
Dimensions in mm
F
GC
250
E
B
100 450
Fig. P1.13
500 675 825
23 Problems 1.14 A couple M of magnitude 1500 N ? m is applied to the crank of
an engine. For the position shown,
determine (a) the force P
required to hold the engine system in equilibrium, (b) the average
P
normal stress in the connecting rod BC, which has a 450-mm2
uniform cross section.
1.15 When the force P reached 8 kN, the wooden specimen shown failed C
in shear along the surface indicated by the dashed line. Determine
the average shearing stress along that surface at the time of failure. 200 mm 15 mm
B
P' P
Fig. P1.15
80 mm
M
60 mm
A
90 mm Steel Wood
1.16 The wooden members A and B are to
be joined by plywood splice Fig. P1.14 plates
that will be fully glued on the surfaces in
contact. As part of the design of the joint,
and knowing that the clearance between the
ends of the members is to be 14 in.,
determine the smallest allowable length L if
the average shearing stress in the glue is not
0.25 in.
A
5.8 kips
to exceed 120 psi.
L
1.17 A load P is applied to a steel rod
supported as shown by an alu minum plate
into which a 0.6-in.-diameter hole has been 4 in.
drilled. Knowing that the shearing stress
B
must not exceed 18 ksi in the
5.8 kips
0.6 in.
in. 14
steel rod and 10 ksi in the aluminum plate,
determine the largest load P that can be
applied to the rod.
1.6 in.
0.4 in.
P
Fig. P1.17
1.18 Two wooden planks, each 22 mm thick
and 160 mm wide, are joined by the glued
mortise joint shown. Knowing that the joint
will fail when the average shearing stress in
the glue reaches 820 kPa, determine the
smallest allowable length d of the cuts if the
joint is to withstand an axial load of
magnitude P 5 7.6 kN.
d
Fig. P1.16
Glue
20 mm
P' P 160 mm 160 mm
20 mm
Fig. P1.18
24 Introduction—Concept of Stress 1.19 The load P applied to a steel rod is distributed to a timber support
d
by an annular
washer. The diameter of the rod is 22 mm and the
allowable outer diameter d of the washer,
knowing that the axial normal stress in the
steel rod is 35 MPa and that the average
bearing stress between the washer and
the timber must not exceed 5 MPa.
22 mm
1.20 The axial force in the column
inner diameter of the washer is 25 mm,
supporting the timber beam shown is P 5
which is slightly larger than the diameter 20 kips. Determine the smallest allowable
of the hole. Determine the smallest
length L of the bearing
exceed 400 psi. P L
Fig. P1.19
aaP
6 in.
P
Fig. P1.20
Fig. P1.21
P 40 kN
plate if the bearing stress in the timber is not to
1.21 An axial load P is supported by a short W8 3 40
column of cross sectional area A 5 11.7 in2 and is
distributed to a concrete founda
tion by a square plate as shown. Knowing that the
average normal stress in the column must not exceed
30 ksi and that the bearing stress on the concrete
foundation must not exceed 3.0 ksi, deter
mine the side a of the plate that will provide the most
Fig. P1.22
economical and safe design.
ing, (b) the size of the footing for which the average
1.22 A 40-kN axial load is applied to a short wooden bearing stress in the soil is 145 kPa.
post that is supported by a concrete footing resting
1.23 A 58-in.-diameter steel rod AB is fitted to a round
on undisturbed soil. Determine (a) the maximum
hole near end C of the wooden member CD. For the
bearing stress on the concrete foot
loading shown, determine (a) the maximum average
120 mm 100 mm
normal stress in the wood, (b) the distance b for
which the average shearing stress is 100 psi on the
surfaces indicated by the dashed lines, (c) the
average bearing stress on the wood.
b
b
A
1 in.
750 lb
B
C
b
1500 lb
4 in.
D
750 lb
Fig. P1.23
25 1.24 Knowing that Problems u 5 408 and P 5 9 kN, determine (a) the smallest
allowable diameter of the pin at B
if the average shearing stress
in the pin is not to exceed 120 MPa, (b) the corresponding aver
age bearing stress in member AB at B, (c) the corresponding
average bearing stress in each of the support brackets at B.
P
A
16 mm
750 mm
50 mm
750 mm
B
C
12 mm
A
Fig. P1.24 and P1.25
d
b
t
1.25 Determine the largest load P that can be applied at A when u 5
608, knowing that the average shearing stress in the 10-mmdiameter pin
at B must not exceed 120 MPa and that the average bearing stress
in member AB and in the bracket at B must not exceed 90 MPa.
B
1.26 Link AB, of width b 5 50 mm and thickness t 5 6 mm, is used to
d
support the end of a horizontal beam. Knowing that the average
normal stress in the link is 2140 MPa, and that the average shearing
stress in each of the two pins is 80 MPa, determine (a) the diameter
d of the pins, (b) the average bearing
stress in the link.
4 in. 7 in.
1.27 For the assembly and loading of
Prob. 1.7, determine (a) the average
shearing stress in the pin at B, (b) the
average bearing stress at B in member
BD, (c) the average bearing stress at B in
member ABC, knowing that this member
has a 10 3 50-mm uniform rectangular
cross section.
1.28 The hydraulic cylinder CF, which
partially controls the position of rod DE,
has been locked in the position shown.
Member BD is 5
8 in. thick and is connected to the vertical
rod by a 38-in.-diameter bolt. Determine
(a) the average shearing stress in the bolt,D
(b) the bearing stress at C in member
Fig. P1.26
BD.
8 in.
A
C B 1.8 in.
20 75
E
400 lb
F
Fig. P1.28
26 Introduction—Concept of Stress 1.11 STRESS ON AN OBLIQUE PLANE UNDER
AXIAL LOADING
P'
P
P'
P
In the preceding sections, axial forces exerted on a two-force
shearing stresses on the
other, was because
stresses were being
determined only on
planes perpendicular to
member (Fig. 1.26a)
the axis of the member
were found to cause
normal stresses in that or connection. As you will
see in this section, axial
member (Fig. 1.26b),
while transverse forces forces cause both normal
and shearing stresses on
exerted on bolts and
planes which are not
pins (Fig. 1.27a) were
found to cause shearing perpendicular to the axis
of the member. Similarly,
stresses in those
connections (Fig. 1.27b). trans
verse forces exerted on a
The reason such a
bolt or a pin cause both
relation was observed
between axial forces and normal and shearing
normal stresses on one stresses on planes which
are not perpendicular to
hand, and trans
the axis of the bolt or
verse forces and
pin.
P'
(a) (b)
Fig. 1.26 Axial forces.
P P P'
P' P'
body that the
distributed forces acting
on the section must be
Consider the two-force equivalent to the force
P.
member of Fig. 1.26,
Resolving P into
which is subjected to
components
F and V,
axial forces P and P9. If
respectively normal and
we pass a section
forming an angle u with tangential to the section
(Fig. 1.28c), we have
a normal plane (Fig.
1.28a) and draw the
free-body diagram of
the portion of member
located to the left of that
section (Fig. 1.28b), we
find from the equilibrium
conditions of the free
Fig. 1.27 Transverse forces.
P
P' P'
P
(a) (b)
A
(a) (b)
sin u (1.12)
F 5 P cos u
V5P
forces distributed
over the section, and the force V the resultant
of shearing forces (Fig. 1.28d). The average
values of the corresponding normal and
A0
P
P'
F
The force F represents the resultant of normal
shearing stresses are obtained
by dividing, respectively, F and
V by
V
P'
the area Au of the section:
Fig. 1.28
(c) (d)
F
s 5 Au
V
t 5 Au(1.13)
the area of a section perpendicular to the axis of the member, we 27
Substituting for F and V from
(1.12) into (1.13), and observing
from Fig. 1.28c that A0 5 Au cos
u, or Au 5 A0ycos u, where A0
denotes
1.12 Stress Under General Loading Conditions;
obtain
A0ycos u
t 5 P sin u
Components of Stress
or
s 5 P cos u
A0ycos u
P
s 5 A0 cos2 u
P
t 5 A0 sin u cos u (1.14)
We note from the first of Eqs. (1.14) that the
normal stress s is maximum when u 5 0, i.e.,
when the plane of the section is per pendicular
to the axis of the member, and that it
approaches zero as u approaches 908. We
check that the value of s when u 5 0 is
P
sm 5 A0(1.15)
as we found earlier in Sec. 1.3. The second of
Eqs. (1.14) shows that the shearing stress t is
zero for u 5 0 and u 5 908, and that for u 5 458
it reaches its maximum value
P
P
tm 5 A0 sin 45° cos 45° 5 2A0(1.16)
The first of Eqs. (1.14) indicates that, when u 5
458, the normal stress s9 is also equal to
Py2A0:
P
s¿ 5 A0 cos2 45° 5 P
2A0(1.17)
The results obtained in Eqs. (1.15), (1.16), and
(1.17) are shown graphically in Fig. 1.29. We
note that the same loading may produce either
a normal stress sm 5 PyA0 and no shearing
stress (Fig. 1.29b), or a normal and a shearing
stress of the same magni
tude s9 5 tm 5 Py2A0 (Fig. 1.29 c and d),
depending upon the orientation of the section.
1.12 STRESS UNDER GENERAL
LOADING CONDITIONS;
COMPONENTS OF STRESS
The examples of the previous sections were
limited to members under axial loading and
connections under transverse loading. Most
structural members and machine components
are under more involved loading conditions.
Consider a body subjected to several loads P1,
P2, etc. (Fig. 1.30). To understand the stress
condition created by these loads at some point
Q within the body, we shall first pass a section
through Q, using a plane parallel to the yz
plane. The portion of the body to the left of the
section is subjected to some of the original
loads, and to normal and shearing forces
distributed over the section. We shall denote by
DFx and DVx, respectively, the normal and the
shearing
–45°
(d) Stresses for
Fig. 1.29
P'
P
(a) Axial loading
P2 y
P3
m
P/A0
(b) Stresses for 0
' P/2A0
m
1
P4
x
P/2A0
(c) Stresses for
m
P
z
45°
Fig. 1.30
y
P/2A0
28 Introduction—Concept of Stress
y
P2 P2 Vx y
' P/2A0
Vx
A
V
z
Fx
x
Fx
x
P1
z
QQ
x
z
(a) (b)
P1
y
the nor mal force DFx has a well-defined
direction, the shearing force DVx
xz
xy
may have any direction in the plane of the
section. We therefore
Fig. 1.31
Q
x
resolve DVx into two component forces,
DVxy and DVxz, in directions parallel to
forces acting on a small area DA
the y and z axes, respectively (Fig.
surrounding point Q (Fig. 1.31a). Note
that the superscript x is used to indicate 1.31b). Dividing now the magnitude of
each force by the area DA, and letting DA
that the forces DFx and DVx act on a
surface perpendicular to the x axis. While approach zero, we define the three
stress components shown in Fig. 1.32:
sx 5 lim ¢AS0
(1.18)
¢Vyx
x
z
¢Fx ¢A
txy 5 lim ¢AS0
Fig. 1.32
¢A
txz 5 lim
¢Vzx ¢A
¢AS0
otherwise. Similarly, corresponding
the shearing stress stress components,
components txy and but since the section
txz are positive if the in Fig. 1.33 now
faces the negative x
corresponding
x
axis, a positive sign
We
note
that
the
first
arrows
y
subscript in sx, txy, point, respectively, in for sx will indicate
the positive y and z
and txz is used to
that the corre
directions. The
indicate that the
above analysis may sponding arrow
stresses under
also be carried out points in the negative
consideration are
exerted on a surface by considering the x direction. Similarly,
per pendicular to the portion of body
positive signs for txy
x axis. The second located to the right of and txz will indicate
subscript in txy and txz the vertical plane
that the
through Q (Fig.
identifies the
corresponding
1.33). The same
direction of the
arrows point,
magnitudes, but
component. The
respectively, in the
normal stress sx is opposite directions, negative y and z
z
are obtained for the directions, as shown
positive if the
xz
shearing in
corresponding arrow normal and
x
x
points in the positive forces DF , DVy ,
x
x direction, i.e., if the and DVz . Therefore,
x
body is in tension,
the same values are
and negative
also obtained for the
xy
Fig. 1.33.
Fig. 1.33
Passing a section through Q parallel to the zx plane, we define 29 1.12 Stress Under General Loading Conditions;
Q
in the same manner the stress components, sy, Q, we shall consider a small cube of side a
tyz, and tyx. Finally, a section through Q parallel centered at Q and the stresses exerted on each
to the xy plane yields the components y
of the six faces of the cube (Fig. 1.34).
Components of Stress
sz, tzx, and tzy.
y
To facilitate the visualization of the stress
condition at point a
yx
yz
xy
zy
The stress components shown the shearing stress exerted on
a
in the figure are sx, sy, and sz, the face perpendicular to the x
axis, while tyx represents the x
aQ
component of the shearing
z
x
stress exerted on the face
which represent the normal
perpendicular to the y axis.
Fig. 1.34
stress on faces respectively
Note that only three faces of
perpen dicular to the x, y, and z the cube are actually visible in
y
axes, and the six shearing
xz zx
Fig. 1.34, and that equal and
stress compo nents txy, txz, etc. opposite stress components act
We recall that, according to the on the hidden faces. While the
definition of the shearing
z
stress components, txy
represents the y component of
are obtained by multiplying the additional
corresponding stress
components by the area DA
of each face. We first write the
stresses acting on the faces of following three equilibrium
the cube differ slightly from the equations:
x
y
A
A
oFx 5 0
oFy 5 0
oFz 5 0
(1.19)
stresses at Q, the error
z
involved is small and vanishes Since forces equal and
as side a of the cube
opposite to the forces actually Fig. 1.35
approaches zero.
shown in Fig. 1.35 are acting
A
yz
xy A zy A
Important relations among the on the hidden faces of the
Q
shearing stress components
cube, it is clear that Eqs. (1.19) A
A
A
will now be derived. Let us
are satisfied. Considering now x z zx A xz
consider the free-body diagram the moments of the forces
of the small cube centered at about axes x9, y9, and z9
x
point Q (Fig. 1.35). The normal drawn from Q in directions
and shearing forces acting on respectively parallel to the x, y,
the various faces of the cube and z axes, we write the three
yx
oMx¿ 5 0
5 0 (1.20)
oMy¿ 5 0
equations
Using a projection on the
x9y9 plane (Fig. 1.36), we
note that the only forces with
moments about the z axis
xA
different from zero are the
yA
shearing forces. These
forces form two couples, one yx A
xy A
of counter clockwise
therefore,
DA)a,
(positive) moment (txy
1loMz 5 0: (txy DA)a 2 (tyx
the other of clockwise
DA)a 5 0
(nega
tive) moment 2(tyx
DA)a. The last of the three
Eqs. (1.20) yields,
from which we conclude that
oMz¿
The relation obtained shows
that the y component of the
shearing stress exerted on a
face perpendicular to the x
axis is equal to the x
z'
xy
A
x
A
a
yx
y
A
A
Fig. 1.36
x'
y'
txy 5 tyx (1.21)
30 Introduction—Concept of Stress component of the shearing stress exerted on a face perpendicular to the y
P
P'
Q
axis. From the remaining two equations (1.20), we derive in a
the condition of
similar manner the
We conclude from stress at a given
relations
Eqs. (1.21) and
(1.22) that only six
stress
tyz 5 tzy tzx 5 txz
components are
(1.22)
required to define
(a) (b)
Fig. 1.37
y
P'
tzx. We also note that, at a given point, shear
cannot take place in one plane only; an equal
shearing stress must be exerted on another
plane perpendicular to the first one. For
example, considering again the bolt of Fig. 1.27
and a small cube at the center Q of the bolt (Fig.
1.37a), we find that shearing stresses of equal
magnitude must be exerted on the two horizontal
faces of
point Q, instead of nine as originally assumed.
These six components are sx, sy, sz, txy, tyz, and
P
that are perpendicular to the
x
P
the cube and on the two faces
forces x
under
axial
loading.
If Sec. 1.11, we find that
z
Before concluding our we con sider a small the conditions of
(a)
stress in the member
discussion of stress cube with faces
x
respectively parallel to may be described as
components, let us
the faces of the
shown in Fig.
consider again the
P and P9 (Fig. 1.37b). case of a member
member and recall the
A
results obtained in
'
sx exerted on the faces of
1.38a; the only stresses are normal stresses
However, if the small
P P'
the cube which are
'
45
perpendicular to the x axis.
P
m m
2A
'
orientation of the sections
considered in Fig. 1.29c and
2A
P
'
(b)
cube is rotated by 458 about
the z axis so that its new
orientation matches the
Fig. 1.38
are exerted on four faces of
d, we conclude that normal
and shearing stresses of equal the cube (Fig. 1.38b). We thus
observe
magnitude
structures and machines that will safely and
economically perform a specified function.
that the same loading condition may lead to
different interpretations of the stress situation at
a given point, depending upon the orientation of a. Determination of the Ultimate Strength of a
the element considered. More will be said about Material. An important element to be considered
by a designer is how the material that has been
this in Chap 7.
selected will behave under a load. For a given
material, this is determined by performing
1.13 DESIGN CONSIDERATIONS
specific tests on prepared samples of the
material. For example, a test specimen of steel
In the preceding sections you learned to
determine the stresses in rods, bolts, and pins may be pre
pared and placed in a laboratory testing machine
under simple loading conditions. In later chap
ters you will learn to determine stresses in more to be subjected to a known centric axial tensile
complex situations. In engineering applications, force, as described in Sec. 2.3. As the
magnitude of the force is increased, various
however, the determination of stresses is
seldom an end in itself. Rather, the knowledge of changes in the specimen are measured, for
stresses is used by engineers to assist in their example, changes in its length and its diameter.
most important task, namely, the design of
Eventually the largest force which may be applied to the specimen is 31
1.13 Design Considerations
reached, and the specimen either breaks or begins to carry less load.
This largest force is called the ultimate load for the test specimen and
is denoted by PU. Since the applied load is centric, we may divide the
ultimate load by the original cross-sectional area of the rod to obtain
the ultimate normal stress of the material
factor of safety.† We have
used. This stress, also known as the ultimate
strength in tension of the material, is
Factor of safety 5 F.S. 5 ultimate load
sU 5 PU
allowable load (1.24)
A (1.23)
An alternative definition of the factor of safety
is based on the use of stresses:
Several test procedures are available to
determine the ultimate shearing stress, or Factor of safety 5 F.S. 5 ultimate stress
ultimate strength in shear, of a material. The
one most commonly used involves the
allowable stress (1.25)
twisting of a circular tube (Sec. 3.5). A more
direct, if less accurate, procedure consists in The two expressions given for the factor of
safety in Eqs. (1.24) and (1.25) are identical
clamping a rectangular or round bar in a
when a linear relationship exists between the
shear tool (Fig. 1.39) and applying an
increasing load P until the ultimate load PU load and the stress. In most engineering
for single shear is obtained. If the free end of applications, however, this rela tionship
the specimen rests on both of the hard ened ceases to be linear as the load approaches
dies (Fig. 1.40), the ultimate load for double its ultimate value, and the factor of safety
shear is obtained. In either case, the ultimateobtained from Eq. (1.25) does not provide a
shearing stress tU is obtained by dividing the
ultimate load by the total area over which
†In some fields of engineering, notably aeronautical
shear has taken place. We recall that, in the engineering, the margin of safety is used in place of the
case of single shear, this area is the cross- factor of safety. The margin of safety is defined as the
sectional area A of the specimen, while in factor of safety minus one; that is, margin of safety 5 F.S. 2
double shear it is equal to twice the cross- 1.00.
P
sectional area.
b. Allowable Load and Allowable Stress;
Factor of Safety. The maximum load that a
structural member or a machine component
will be allowed to carry under normal
conditions of utilization is consider ably
smaller than the ultimate load. This smaller
Fig. 1.39 Single shear test. P
load is referred to as the allowable load and,
sometimes, as the working load or design
load. Thus, only a fraction of the ultimateload capacity of the member is utilized when
the allowable load is applied. The remaining
portion of the load-carrying capacity of the
member is kept in reserve to assure its safe
performance. The ratio of the ultimate load to
Fig. 1.40 Double shear test.
the allowable load is used to define the
32 Introduction—Concept of Stress true assessment of the safety of a given design. Nevertheless, the allowablestress method of design, based on the use of Eq. (1.25), is
widely used.
c. Selection of an Appropriate Factor of Safety. The selec
tion of the factor of safety to be used for various applications is one
of the most important engineering tasks. On the one hand, if a factor
of safety is chosen too small, the possibility of failure becomes unac
ceptably large; on the other hand, if a factor of safety is chosen
unnecessarily large, the result is an uneconomical or nonfunctional
design. The choice of the factor of safety that is appropriate for a
given design application requires engineering judgment based on
many considerations, such as the following:
1. Variations that may occur in the properties of the member
under consideration. The composition, strength, and dimen
sions of the member are all subject to small variations during
manufacture. In addition, material properties may be altered
and residual stresses introduced through heating or deforma
tion that may occur during manufacture, storage, transporta
tion, or construction.
2. The number of loadings that may be expected during the life of
the structure or machine. For most materials the ultimate stress
decreases as the number of load applications is increased. This
phenomenon is known as fatigue and, if ignored, may result in
sudden failure (see Sec. 2.7).
3. The type of loadings that are planned for in the design, or that
may occur in the future. Very few loadings are known with
complete accuracy—most design loadings are engineering esti
mates. In addition, future alterations or changes in usage may
introduce changes in the actual loading. Larger factors of safety
are also required for dynamic, cyclic, or impulsive loadings.
4. The type of failure that may occur. Brittle materials fail sud
denly, usually with no prior indication that collapse is immi
nent. On the other hand, ductile materials, such as structural
steel, normally undergo a substantial deformation called yield
ing before failing, thus providing a warning that overloading
exists. However, most buckling or stability failures are sudden,
whether the material is brittle or not. When the possibility of
sudden failure exists, a larger factor of safety should be used
than when failure is preceded by obvious warning signs.
5. Uncertainty due to methods of analysis. All design methods are
based on certain simplifying assumptions which result in calcu
lated stresses being approximations of actual stresses.
6. Deterioration that may occur in the future because of poor main
tenance or because of unpreventable natural causes. A larger fac
tor of safety is necessary in locations where conditions such as
corrosion and decay are difficult to control or even to discover.
7. The importance of a given member to the integrity of the whole
structure. Bracing and secondary members may in many cases
be designed with a factor of safety lower than that used for
primary members.
In addition to the these considerations, there is the additional 33
1.13 Design Considerations
consideration concerning the risk to life and property that a failure
would produce. Where a failure would produce no risk to life and
only minimal risk to property, the use of a smaller factor of safety
can be considered. Finally, there is the practical consideration that,
unless a careful design with a nonexcessive factor of safety is used,
a structure or machine might not perform its design function. For
example, high factors of safety may have an unacceptable effect on
the weight of an aircraft.
For the majority of structural and machine applications, factors
of safety are specified by design specifications or building codes writ
ten by committees of experienced engineers working with profes
sional societies, with industries, or with federal, state, or city agencies.
Examples of such design specifications and building codes are
1. Steel: American Institute of Steel Construction, Specification
for Structural Steel Buildings
2. Concrete: American Concrete Institute, Building Code Require
ment for Structural Concrete
3. Timber: American Forest and Paper Association, National
Design Specification for Wood Construction
4. Highway bridges: American Association of State Highway Offi
cials, Standard Specifications for Highway Bridges
*d. Load and Resistance Factor Design. As we saw previously,
the allowable-stress method requires that all the uncertainties associ
ated with the design of a structure or machine element be grouped
into a single factor of safety. An alternative method of design, which
is gaining acceptance chiefly among structural engineers, makes it pos
sible through the use of three different factors to distinguish between
the uncertainties associated with the structure itself and those associ
ated with the load it is designed to support. This method, referred
to as Load and Resistance Factor Design (LRFD), further allows the
designer to distinguish between uncertainties associated with the live
load, PL, that is, with the load to be supported by the structure, and
the dead load, PD, that is, with the weight of the portion of structure
contributing to the total load.
When this method of design is used, the ultimate load, PU, of
the structure, that is, the load at which the structure ceases to be
useful, should first be determined. The proposed design is then
acceptable if the following inequality is satisfied:
gDPD 1 gLPL # fPU (1.26)
The coefficient f is referred to as the resistance factor; it accounts
for the uncertainties associated with the structure itself and will
normally be less than 1. The coefficients gD and gL are referred to
as the load factors; they account for the uncertainties associated,
respectively, with the dead and live load and will normally be greater
than 1, with gL generally larger than gD. While a few examples or
assigned problems using LRFD are included in this chapter and in
Chaps. 5 and 10, the allowable-stress method of design will be used
in this text.
P
0.6 m
dAB
A
B
determine the
diameter of the
rod for which
the factor of
safety with
respect to
failure will be
B
3.3. (b) The pin
at C is to be
made of a steel
ttP
50 kN 15 kN
0.3 m 0.3 m
SAMPLE
PROBLEM
1.3
Two forces are
applied to the
bracket BCD as
shown. (a) Knowing that the of a steel having an ultimate
control rod AB is to be madenormal stress of 600 MPa,
D
C
having an ultimate shearing
stress of 350 MPa.
Determine the diameter of
the pin C for which the
factor of safety with respect
to shear will also be 3.3. (c)
Determine the required
thickness of the bracket
supports at C
knowing that the allowable
bearing stress of the steel
used is 300 MPa.
SOLUTION
Free Body: Entire Bracket.
The reaction at C is
represented by its com
50 kN 15 kN
1 l oMC 5 0: P(0.6 m) 2 (50 kN)(0.3 m) 2 (15 kN)(0.6
0.6 m C
m) 5 0 P 5 40 kN oFx 5 0: Cx
76.3 kN
ponents Cx and Cy.
oFy 5 0: Cy 5 65 kN
Cx
C
F2
a. Control Rod AB. Since the
factor of safety is to be 3.3, the
Cy
dC
D
0.3 m 0.3 m
5 40 k
C 5 2C2x 1 C2y
5
allow
able
stress
is
sall 5
sU
F.S.5
600
MPa
3.3 5
181.8
MPa
For P
5 40
kN the cross-sectional area
required is
P
Areq 5 sall5 40 kN
181.8 MPa 5 220 3 1026 m2
p
Areq 5 4dAB
2 5 220 3 1026 m2 d
AB 5 16.74 mm
◀
b. Shear in Pin C. For a factor of
safety of 3.3, we have tall 5 tU
F.S.5 350 MPa
F1F1 F2
C
3.3 5 106.1 MPa
12
Since the pin is in double shear, we
write
tall5 176.3 kN2y2
Areq 5
Cy2
106.1 MPa 5 360
mm2
dC 5 22 mm ◀ The next larger size pin
available is of 22-mm diameter and
should be used.
c. Bearing at C. Using d 5 22 mm, the
nominal bearing area of each bracket
is 22t. Since the force carried by each
bracket is Cy2 and the allow able
bearing stress is 300 MPa, we write
sall5 176.3 kN2y2
C12
t C12
Areq 5
p
4dC2
5 360
mm2
dC 5
21.4
mm
Use:
d 22 mm
300 MPa 5 127.2 mm2
Areq 5 Cy2
Thus 22t 5 127.2 t 5
5.78 mm Use: t 5 6 mm
◀
34
C
B
B
A
C
D 8 in.
6 in.
C
SAMPLE PROBLEM 1.4
SOLUTION
The rigid beam BCD is attached by
bolts to a control rod at B, to a
hydraulic cylinder at C, and to a fixed
support at D. The diameters of the
bolts used are: dB 5 dD 5 38 in., dC 5 12
in. Each bolt acts in double shear and
is made from a steel for which the
The factor of safety with respect to
failure must be 3.0 or more in each of
ultimate shearing stress is tU
D
5 40 ksi.
The con
trol rod AB has a diameter
dA 5 716 in. and is made of a steel for
which the ultimate tensile stress is sU
5 60 ksi. If the minimum factor of
safety is to be 3.0 for the entire unit,
determine the largest upward force
which may be applied by the
hydraulic cylinder at C.
the three bolts and in the control rod.
These four independent criteria will
be considered separately.
B D 6 in. 8
in.
Free
Body:
Beam
BCD. We
first
determine
the force
at C in
terms of
the force
at B and
in terms
of the
force
at D.
1l oMD
5 0:
B114
in.2 2
C18
in.2 5
0C5
1.750B
(1) 1l
oMB 5
0:
2D114
in.2 1
C16
in.2 5
0C5
2.33D
(2)
Control Rod. For a factor of safety of 3.0 we have
sall 5 sU
F.S.5 60 ksi
3.0 5 20 ksi
The allowable force in the control rod is
Using Eq. (1) we find the largest permitted value of C:
3
8 in.
C 5 1.750B 5 1.75013.01 kips2 C 5 5.27 kips ◀
B 5 sall1A2 5 120
ksi214p
1
7 in.22 5
16
3.01 kips
F1
Bolt at B. tall 5 tUyF.S. 5 (40 ksi)y3 5 13.33 ksi. Since
the bolt is in
double shear, the allowable magnitude of the force B
exerted on the bolt is
F1
B 2F1
B 5 2F1 5 21tall A2 5 2113.33 ksi2114 p2138 in.22 5 2.94
kips
From Eq. (1): C 5 1.750B 5 1.75012.94 kips2 C 5
B
5.15 kips ◀
C
is the same as bolt B, the Summary. We have
allowable force is D 5 B 5found separately four
2.94 kips. From Eq. (2): maximum allowable
values of the force C. In
order to satisfy all these
criteria we must choose
C 5 2.33D 5 2.3312.94
the smallest value,
kips2 C 5 6.85 kips ◀
namely: C 5 5.15 kips ◀
C 2F2
1
2 in.
Bolt at C. We again have
tall 5 13.33 ksi and write C
5 2F2 5 21tall A2 5
2113.33 ksi2114 p2112
F2
◀
2
Bolt at D. Since this bolt in.2 C 5 5.23 kips
F2
35
PROBLEMS
shown. Determine the normal and shearing
stresses in the glued splice.
P
5.0 in.
3.0 in.
1.30 Two wooden members of uniform cross
section are joined by the simple scarf splice
shown. Knowing that the maximum allowable
tensile stress in the glued splice is 75 psi,
determine (a) the largest load P that can be
safely supported, (b) the corresponding
shearing stress in the splice.
1.31 Two wooden members of uniform
rectangular cross section are joined by the
simple glued scarf splice shown. Knowing that
60
P 5 11 kN, determine the normal and shearing
1.29 The 1.4-kip load P is supported by two
wooden members of uni form cross section that stresses in the glued splice.
are joined by the simple glued scarf splice
P
45
75 mm
150 mm
P'
Fig. P1.29 and P1.30
P'
1.32 Two wooden members of uniform rectangular
cross section are joined by the simple glued scarf
splice shown. Knowing that the maximum allowable
shearing stress in the glued splice is 620 kPa,
determine (a) the largest load P that can be safely
applied, (b) the corresponding tensile stress in the
splice.
P
1.33 A steel pipe of 12-in. outer diameter is fabricated
from 14-in.-thick plate by welding along a helix that
forms an angle of 258 with a
in. 14
Weld
plane perpendicular to the axis of the pipe. Knowing
that the maxi mum allowable normal and shearing
stresses in the directions respectively normal and
tangential to the weld are s 5 12 ksi and t 5 7.2 ksi,
determine the magnitude P of the largest axial force
that can be applied to the pipe.
25
Fig. P1.33 and P1.34
Fig. P1.31 and P1.32
1.34 A steel pipe of 12-in. outer diameter is fabricated
from 14-in.-thick plate by welding along a helix that
forms an angle of 258 with a plane perpendicular to
the axis of the pipe. Knowing that a 66 kip axial force
P is applied to the pipe, determine the normal and
shearing stresses in directions respectively normal
and tangential to the weld.
36
37 Problems 1.35 A 1060-kN load P is applied to the granite block shown. Determine
the resulting maximum value of (a) the normal
stress, (b) the shear ing stress. Specify the
orientation of that plane on which each of these
maximum values occurs.
1.36 A centric load P is applied to the granite block
shown. Knowing that the resulting maximum value
of the shearing stress in the block is 18 MPa,
determine (a) the magnitude of P, (b) the orienta
tion of the surface on which the maximum shearing
stress occurs, (c) the normal stress exerted on that
140 mm
surface, (d) the maximum value of the normal
stress in the block.
140 mm
a 16-kN load P?
1.37 Link BC is 6 mm thick, has a width w 5 25
mm, and is made of a steel with a 480-MPa
1.38 Link BC is 6 mm thick and is made of a steel
ultimate strength in tension. What is the safety
with a 450-MPa ultimate strength in tension. What
factor used if the structure shown was designed to should be its width w if the structure shown is being
support
designed to support a 20-kN load P with a factor of
safety of 3?
P
1.39 A 34-in.-diameter rod made of the same
material as rods AC and AD in the truss shown was
tested to failure and an ultimate load of 29 kips was 90
recorded. Using a factor of safety of 3.0, determine 480 mm
the required diameter (a) of rod AC, (b) of rod AD.
C
D
A
Fig. P1.35 and P1.36 600 mm
P
ABw
5 ft
D
10 ft 10 ft
Fig. P1.37 and P1.38
BC
10 kips 10 kips
Fig. P1.39 and P1.40
1.40 In the truss shown, members AC and AD consist of rods made of
the same metal alloy. Knowing that AC is of 1-in. diameter and
that the ultimate load for that rod is 75 kips, determine (a) the
factor of safety for AC, (b) the required diameter of AD if it is
desired that both rods have the same factor of safety.
1.41 Link AB is to be made of a steel for which the ultimate normal stress
is 450 MPa. Determine the cross-sectional area of AB for which
the factor of safety will be 3.50. Assume that the link will be
adequately reinforced around the pins at A and B.
8 kN/m
A
C DE
35
20 kN
B
0.4 m 0.4 m
0.4 m
Fig. P1.41
38 Introduction—Concept of Stress 1.42 A steel loop ABCD of length 1.2 m and of 10-mm diameter is
placed as shown
around a 24-mm-diameter aluminum rod AC.
240 mm
steel used for the loop determine the larg
Cables BE and DF, each and the cables is 480
est load Q that can be
of 12-mm diameter, are MPa and that the
applied if an overall
used to apply the load ultimate strength of the factor of safety of 3 is
Q. Knowing that the
aluminum used for the desired.
ultimate strength of the rod is 260 MPa,
240 mm
E
QB
180 mm
ultimate shearing stress in the glue
is 360 psi and the clearance
between the members is 14 in.
10 mm
Determine the required length L of
1.43 Two wooden members shown, each splice if a factor of safety of
2.75 is to be achieved.
which support a 3.6-kip load, are
joined by plywood splices fully glued
on the surfaces in contact. The
24 mm
A
180 mm
C
D
F
12 mm
Fig. P1.42
Q'
3.6 kips
5.0 in.
Fig. P1.43
in. 14
L
3.6 kips
shown. Knowing that the ultimate
shearing stress of the bonding
between the surfaces is 130 psi,
determine the factor of safety with
respect to shear when P 5 325 lb.
P'
P
in. 58 in. 34 in.2 14
1.45 A load P is supported as
shown by a steel pin that has been
inserted in a short wooden member
1.44 Two plates, each 18-in. thick, hanging from the ceiling. The
are used to splice a plastic strip as ultimate
in. 14
Fig. P1.44
strength of the wood used is 60 MPa in tension and
7.5 MPa in shear, while the ultimate strength of the
steel is 145 MPa in shear. Knowing that b 5 40 mm, c
5 55 mm, and d 5 12 mm, determine the load P if an
overall factor of safety of 3.2 is desired.
d
1
P
2
1
Fig. P1.45
40 mm
c
b
2P
39 1.46 For the support of Prob. 1.45, knowing that the diameter of Problems the pin is d 5 16 mm and that the
magnitude of the load is
P 5 20 kN, determine (a) the factor of safety for the pin, (b) the
required values of b and c if the factor of safety for the wooden
member is the same as that found in part a for the pin.
1.47 Three steel bolts are to be used to attach the steel plate shown to
a wooden beam. Knowing that the plate will support a 110-kN
load, that the ultimate shearing stress for the steel used is 360 MPa,
and that a factor of safety of 3.35 is desired, determine the required
diameter of the bolts.
1.48 Three 18-mm-diameter steel bolts are to be used to attach the steel
plate shown to a wooden beam. Knowing that the plate will support
a 110-kN load and that the ultimate shearing that width should be embedded in the concrete
slab. (Neglect the normal stresses between the
stress for the steel used is 360 MPa,
determine the factor of safety for this design. concrete and the bottom edge of the plate.)
1.49 A steel plate 516 in. thick is embedded in a
horizontal concrete slab and is used to anchor P
a high-strength vertical cable as shown. The
diameter of the hole in the plate is 34 in., the
ultimate strength of the steel used is 36 ksi,
and the ultimate bonding stress between plate
and concrete is 300 psi. Knowing that a factor
of safety of 3.60 is desired when P 5 2.5 kips,
determine (a) the required width a of the plate,
(b) the minimum depth b to which a plate of
5
3
16 in.
110 kN
Fig. P1.47 and P1.48
Fig. P1.49
4 in.
a
b
40
1.50 Determine the factor of safety for the cable anchor in Prob. 1.49
when P 5 3 kips, knowing that a 5 2 in. and b 5 7.5 in.
Introduction—Concept of Stress 1.51 In the steel structure shown, a 6-mm-diameter pin is used at C and 10-mmdiameter pins are used at B and D. The ultimate
shearing stress is 150 MPa at all connections, and the ultimate
normal stress is 400 MPa in link BD. Knowing that a factor
of safety of 3.0 is desired, determine the largest load P that can
be applied at A. Note that link BD is not reinforced around the
pin holes.
Front view
D
18 mm
DB
A
6 mm
160 mm 120 mm
P
A
Side view
B
C
B
C
Top view
Fig. P1.51
1.52 Solve Prob. 1.51, assuming that the structure has been redesigned
to use 12-mm-diameter pins at B and D and no other change
has been made.
1.53 Each of the two vertical links CF connecting the two horizontal
mem bers AD and EG has a uniform rectangular cross section
1 in. thick and 1 in. wide, and is made of a steel with an
4
ultimate strength in tension of 60 ksi. The pins at C and F each
have a 12-in. diameter and are made of a steel with an ultimate
strength in shear of 25 ksi. Determine the overall factor of
safety for the links CF and the pins connecting them to the
horizontal members.
10 in.
16 in.
10 in.
A
B
D
E
Fig. P1.53
F
G 2 kips
C
41 1.54 Problems Solve Prob. 1.53, assuming that the pins at C and F have been
diameter.
1.55 In the structure shown, an 8-mm-diameter pin is used at A, and
12-mm-diameter pins are used at B and D. Knowing that the ulti
mate shearing stress is 100 MPa at all connections and that the
ultimate normal stress is 250 MPa in each of the two links joining
B and D, determine the allowable load P if an overall factor of
safety of 3.0 is desired.
Top view
200 mm 180 mm
12 mm
8 mm
ABC
replaced by pins with a 34-in.
B
C
A
B
20 mm 8 mm
8 mm
P
DD
Side view
Front view
Fig. P1.55
12 mm
PP
platform support ing two
safety for the selected
window washers. The
cables?
platform weighs 160 lb and
each of the window
1.56 In an alternative
design for the structure of washers is assumed to
Prob. 1.55, a pin of 10-mm weigh 195 lb with equip
diameter is to be used at A. ment. Since these workers
are free to move on the
Assuming that all other
platform, 75% of their total
speci fications remain
unchanged, determine the weight and the weight of
their equipment will be used
allowable load P if an
overall factor of safety of 3.0 as the design live load of
each cable. (a) Assuming a
is desired.
Fig. P1.57 1.8 m
resis tance factor f 5 0.85
and load factors gD 5 1.2
*1.57 The Load and
Resistance Factor Design and gL 5 1.5, determine the
required minimum ultimate
method is to be used to
load of one cable. (b) What
select the two cables that
is the conventional factor of C
will raise and lower a
placed on the platform. (b) What is the
corresponding conventional factor of safety for rod
BC?
AB
*1.58 A 40-kg platform is attached to the end B of
a 50-kg wooden beam AB, which is supported as 2.4 m
shown by a pin at A and by a slender steel rod BC
with a 12-kN ultimate load. (a) Using the Load and
Resistance Factor Design method with a resistance
factor f 5 0.90 and load factors gD 5 1.25 and gL
1.6, deter
5
mine the largest load that can be safely
Fig. P1.58
REVIEW AND
SUMMARY
This chapter was devoted to the concept of stress and to an introduc
tion to the methods used for the analysis and design of machines
and load-bearing structures.
Section 1.2 presented a short review of the methods of statics
and of their application to the determination of the reactions
exerted by its supports on a simple structure consisting of pinconnected members. Emphasis was placed on the use of a freebody diagram
to obtain equilibrium equations which were solved for the unknown
reactions. Free-body diagrams were also used to find the internal
forces in the various members of the structure.
Axial loading. Normal stress P
P
s 5 A (1.5)
Section 1.4 was devoted to a short discussion of
the
two prin cipal tasks of an engineer, namely,
A
the
analysis and the design of structures and
The concept of stress was first introduced in Sec.
machines.
1.3 by consider ing a two-force member under an
axial loading. The normal stress in that member As noted in Sec. 1.5, the value of s obtained
was obtained by dividing the magnitude P of the from Eq. (1.5) represents the average stress
load by the cross-sectional area A of the memberover the section rather than the stress at a
specific point Q of the section. Considering a
(Fig. 1.41). We wrote
small area DA surrounding Q and the magnitude
DF of the force exerted on DA, we defined the
stress at point Q as
s 5 lim ¢Ay0
¢A (1.6)
¢F
P'
However, for the distribution of
Fig. 1.41
In general, the value obtained for the stresses to be uniform in a given
stress s at point Q is different from the section, it is necessary that the line of
value of the average stress given by
action of the loads P and P9 pass
formula (1.5) and is found to vary
through the centroid C of the section.
across the section. However, this varia Such a loading is called a centric axial
tion is small in any section away from loading. In the case of an eccentric
the points of application of the loads. axial loading, the distribution of
In practice, therefore, the distribution of stresses is not uniform. Stresses in
the normal stresses in an axially
members sub jected to an eccentric
loaded member is assumed to be
axial loading will be discussed in Chap
uniform, except in the immediate
4.
vicinity of the points of application of
the loads.
42
When equal and opposite transverse forces P and P9 of magnitude 43
Review and Summary
P are applied to a member AB (Fig. 1.42), shearing stresses t are
created over any section located between the
P
F
points of application of the two forces [Sec 1.6].
tave 5 A 5 A (1.9)
These stresses vary greatly across the section
Transverse forces. Shearing stress P
and their distribution cannot be assumed uniform.
However, dividing the magnitude P—referred to as A C B
the shear in the section—
by the cross-sectional area A, we defined the
average shearing stress over the section:
tave 5
P
A (1.8)
P
Fig. 1.42
Shearing stresses are found in bolts, pins, or rivets
connecting two structural members or machine
components. For example, in the case of bolt CD Single and double shear
(Fig. 1.43), which is in single shear, we wrote
C
while, in the case of bolts EG and
HJ (Fig. 1.44), which are both in
double shear, we had
tave 5
P
A 5 Fy2
F'
F
A 5 2A (1.10)
Fig. 1.43
AF
E
B E' D
1.7]. The bolt CD of Fig. 1.43, for example,
Bolts, pins, and rivets also create stresses in creates stresses on the semicylin
the members they con nect, along the
Bearing stress
bearing surface, or surface of contact [Sec.
EH
drical surface of plate A with which it is in contact (Fig. 1.45). Since
C
K
K'
the distribution of these stresses is quite
complicated, one uses in practice an average
nominal value sb of the stress, called bearing
F' F B A
stress, obtained by
P
L'
dividing the load P by
the area of the rectangle
Fig. 1.44
representing the
D
projection of the bolt on
P
P
sb 5 A 5 td (1.11)
the plate section.
GJ
Denoting by t the
In Sec. 1.8, we applied
thickness of the plate
the concept introduced
and by d the diameter of in the previous
C
the bolt, we wrote
t
L
determined succes shearing stresses in
sections to the
the various pins,
analysis of a simple sively the normal
D
F'
structure consisting stresses in the two and the bearing
members, paying stress at each
of two pin
Fig. 1.45
connected members special attention to connection.
A
d
supporting a given their narrowest
F
sections,
the
load. We
use to write equi
librium equations. These equations will be
solved for unknown forces, from which the
required stresses and deformations can be
computed. Once the answer has been
The method you should use in solving a
obtained, it should be care fully checked.
problem in mechanics of materials was
described in Sec. 1.9. Your solution should
begin with a clear and precise statement of Method of Solution
the problem. You will then draw one or
several free-body diagrams that you will
44 Introduction—Concept of Stress The first part of the chapter ended with a discussion of numeri cal accuracy in
engineering, which stressed the fact that the accuracy
of an answer can never be greater than the accuracy of the given
data [Sec. 1.10].
Stresses on an oblique section
P' P
Fig. 1.46
Stress under general loading
P
s 5 A0 cos2 u
y
P
t 5 A0 sin u cos u (1.14)
We observed from these formulas that the normal
y
for u
In Sec. 1.11, we considered the stresses created
stress is maximum and equal to sm 5 PyA0
on an oblique section in a two-force member
5 0, while the shearing stress is maxi
under axial loading. We found that both nor mal
mum
and shearing stresses occurred in such a
and equal to tm 5 Py2A0 for u 5 458. We also
situation. Denoting by u the angle formed by the noted that t 5 0 when u 5 0, while s 5 Py2A0
section with a normal plane (Fig. 1.46) and by A0 when u 5 458.
the area of a section perpendicular to the axis of
the member, we derived the following
Next, we discussed the state of stress at a point
expressions for the normal stress s and the
Q in a body under the most general loading
shearing stress t on the oblique section:
condition [Sec. 1.12]. Considering a small cube
centered at Q (Fig. 1.47), we denoted by sx the
normal stress exerted on a face of the cube
perpendicular to the x axis, and by txy
and z components of the
cube and observing that txy
a
shearing stress exerted on 5 tyx, tyz 5
yx
the same face of the cube.
yz
xy zy
Repeating this procedure for
the other two faces of the
tzy, and tzx 5 txz, we concluded that six
stress components are
required to define the state of
ultimate strength of the material
stress at a given point Q,
used, as deter
namely, sx, sy, sz, txy, tyz, and mined by a laboratory test on a
tzx.
specimen of that material. The
Section 1.13 was devoted to a ulti mate load should be
discussion of the various
considerably larger than the
concepts used in the design of allowable load, i.e., the load that
engineering structures. The
the member or component will
ultimate load of a given
be allowed to carry under
structural member or machine normal conditions. The ratio of
component is the load at which the ultimate load to the allow
the member or component is
able load is defined as the factor
expected to fail; it is computed of safety:
from the ultimate stress or
and txz, respectively, the y
aQ
x
z
a
z
Fig. 1.47
xz zx
x
Factor of safety
Load and Resistance Factor Design
Factor of safety 5 F.S. 5 ultimate load
allowable load (1.24)
The determination of the factor of safety that should
be used in the design of a given structure depends
upon a number of consider ations, some of which
were listed in this section.
Section 1.13 ended with the discussion of an
alternative approach to design, known as Load and
Resistance Factor Design, which allows the
engineer to distinguish between the uncertainties
associated with the structure and those associated
with the load.
REVIEW
PROBLEMS
1.59 A strain gage located at C on the surface of bone AB indicates
that the average normal stress in the bone is 3.80 MPa when
the bone
is subjected to two 1200-N forces as shown. Assuming the cross
section of the bone at C to be annular and knowing that its outer
diameter is 25 mm, determine the inner diameter of the bone’s
cross section at C.
1200 N
A
C
B
1200 N
Fig. P1.59
1.60 Two horizontal 5-kip forces are applied to pin B of the assembly
shown. Knowing that a pin of 0.8-in. diameter is used at each
connection, determine the maximum value of the average normal
stress (a) in link AB, (b) in link BC.
0.5 in.
B
1.8 in.
A
C
60
45
5 kips
5 kips
Fig. P1.60
0.5 in.
1.8 in.
1.61 For the assembly and loading of Prob. 1.60, determine (a) the
average shearing stress in the pin at C, (b) the average bearing
stress at C in member BC, (c) the average bearing stress at B
in member BC.
45
46 Introduction—Concept of Stress 1.62 In the marine crane shown, link CD is known to have a uniform
cross section of 50
3 150 mm. For the loading shown, determine
the normal stress in the central portion of that link.
15 m 25 m 3 m
B
35 m
80 Mg
C
15 m
D
A
Fig. P1.62
1 in. thick and 9 in. wide, are joined by
2
the dry mortise joint shown. Knowing that the wood used shears
off along its grain when the average shearing stress reaches
1.20 ksi, determine the magnitude P of the axial load that will
cause the joint to fail.
1.63 Two wooden planks, each
in. 58
in. 58
1 in.
2 in.
P' 2 in.
P'
1 in. 9 in.P
Fig. P1.63
P
Fig. P1.64
1.64 Two wooden members of uniform rectangular cross
a
b
47
section of sides a 5 100 mm and b 5 60 mm are joined
by a simple glued joint as shown. Knowing that the
ultimate stresses for the joint are sU 5 1.26 MPa in
tension and tU 5 1.50 MPa in shear and that P 5 6 kN,
determine the factor of safety for the joint when (a) a 5
208, (b) a 5 358, (c) a 5 458. For each of these values
of a, also determine whether the joint will fail in tension
or in shear if P is increased until rupture occurs.
Review Problems 1.65 Member ABC, which is supported by a pin and bracket at C and a cable BD, was designed to
support the 16-kN load P as shown.
Knowing that the ultimate load for cable BD is 100 kN, determine
the factor of safety with respect to cable failure.
40
A
0.6 m
P
C
D
Fig. P1.65
0.8 m
0.4 m
30
B
1.66 The 2000-lb load may be moved along the beam BD to any posi
tion between stops at E and F. Knowing that sall 5 6 ksi for the
steel used in rods AB and CD, determine where the stops
should be placed if the permitted motion of the load is to be as
large as possible.
60 in.
AC
-in. 58 -in. 12
xF
diameter diameter
xE
EF
B
Fig. P1.66
D
2000 lb
x
1.67 Knowing that a force P of magnitude 750 N is applied to the pedal
shown, determine (a) the diameter of the pin at C for which the
average shearing stress in the pin is 40 MPa, (b) the
correspond ing bearing stress in the pedal at C, (c) the
corresponding bearing stress in each support bracket at C.
A
B
9 mm
75 mm
P
300 mm
125 mm
CC
D 5 mm
Fig. P1.67
48 Introduction—Concept of Stress 1.68 A force P is applied as shown to a steel reinforcing bar that has
been embedded
in a block of concrete. Determine the smallest
length L for which the full allowable normal stress in the bar can
be developed. Express the result in terms of the diameter d of the
bar, the allowable normal stress sall in the steel, and the average
allowable bond stress tall
between the concrete and the cylindri
cal surface of the bar. (Neglect the normal stresses
between the
concrete and the end of the bar.)
Ld
P
Fig. P1.68
1.69 The two portions of member AB are glued together along a plane
forming an angle u with the horizontal. Knowing that the ultimate
stress for the glued joint is 2.5 ksi in tension and 1.3 ksi in shear,
determine the range of values of u for which the factor of safety
of the members is at least 3.0.
2.4 kips
A
B
2.0 in.
Fig. P1.69 and
P1.70
1.25 in.
1.70 The two portions of member AB are glued together along a plane
forming an angle u with the horizontal. Knowing that the ultimate
stress for the glued joint is 2.5 ksi in tension and 1.3 ksi in shear,
determine (a) the value of u for which the factor of safety of the
member is maximum, (b) the corresponding value of the factor
of safety. (Hint: Equate the expressions obtained for the factors
of safety with respect to normal stress and shear.)
COMPUTER
PROBLEMS
The following problems are designed to be solved with a computer.
1.C1 A solid steel rod consisting of n cylindrical elements welded
together is subjected to the loading shown. The diameter of element i is
denoted by di and the load
connecting pins C and E, (3)
, the average shearing stress in
applied to its lower end by Pi
pin B, (4) the average shearing
with the magni
tude Pi of this stress in pin C, (5) the
load being assumed positive if average bearing stress at B in
Pi is directed downward as
member ABC, (6) the average
shown and negative otherwise. bearing stress at C in member
(a) Write a computer program ABC. (b) Check your program
that can used with either SI or by comparing the values
U.S. customary units to
obtained for d 5 16 mm with
determine the average stress the answers given for Probs.
in each element of the rod. (b)
1.7 and 1.27. (c) Use this
Use this program to solve
program to find the permissible
Probs. 1.2 and 1.4.
values of the diameter d of the
pins, knowing that the
1.C2 A 20-kN load is applied
allowable values of the
as shown to the horizontal
normal, shearing, and bearing
member ABC. Member ABC
stresses for the steel used are,
has a 10 3 50-mm uniform
respectively, 150 MPa, 90
rectangular cross section and
MPa, and 230 MPa. (d) Solve
is supported by four vertical
part c, assuming that the thick
links, each of 8 3 36-mm
ness of member ABC has
uniform rectangular cross
been reduced from 10 to 8
section. Each of the four pins
mm.
at A, B, C, and D has the same
diam
eter d and is in double shear.
(a) Write a computer program
0.4 m
to calculate for values of d
C
from 10 to 30 mm, using 1-mm
increments, (1) the maxi mum
value of the average normal
Pn
stress in the links connecting
pins B and D, (2) the average
normal stress in the links
Fig. P1.C1
Element n Element 1
P1
0.5 in.
0.5 in.
B
1.8 in.
20 kN
B
1.8 in.
A
0.2 m
A
45
5 kips
5 kips
E
0.25 m
60
D
normal stress in member AB, (2) the average normal
stress
Fig. P1.C2
1.C3 Two horizontal 5-kip forces are applied to pin B
C
of the assembly shown. Each of the three pins at A,
B, and C has the same diameter d and is in double
shear. (a) Write a computer program to calculate for
values of d from 0.50 to 1.50 in., using 0.05-in.
increments, (1) the maximum value of the average Fig. P1.C3
49
50 Introduction—Concept of Stress in member BC, (3) the average shearing stress in pin A, (4) the average
shearing
stress in pin C, (5) the average bearing stress at A in member AB,
(6) the average bearing stress at C in member BC, (7) the average bearing
stress at B in member BC. (b) Check your program by comparing the values
obtained for d 5 0.8 in. with the answers given for Probs. 1.60 and 1.61.
(c) Use this program to find the permissible values of the diameter d of the
pins, knowing that the allowable values of the normal, shearing, and bearing
stresses for the steel used are, respectively, 22 ksi, 13 ksi, and 36 ksi. (d) Solve
part c, assuming that a new design is being investigated in which the thick
ness and width of the two members are changed, respectively, from 0.5 to
0.3 in. and from 1.8 to 2.4 in.
1.C4 A 4-kip force P forming an angle a with the vertical is applied
as shown to member ABC, which is supported by a pin and bracket at C
and by a cable BD forming an angle b with the horizontal. (a) Knowing
that the ultimate load of the cable is 25 kips, write a computer program
to construct a table of the values of the factor of safety of the cable for
values of a and b from 0 to 458, using increments in a and b correspond
ing to 0.1 increments in tan a and tan b. (b) Check that for any given
value of a, the maximum value of the factor of safety is obtained for b 5
38.668 and explain why. (c) Determine the smallest possible value of the
factor of safety for b 5 38.668, as well as the corresponding value of a,
and explain the result obtained.
D
P
b
A B
15 in.
C
18 in. 12 in.
P
Fig. P1.C4
a
a simple glued scarf splice. (a) Denoting
by sU and tU, respectively, the ultimate
strength of the joint in tension and in
shear, write a computer program which, for
given values of a, b, P, sU and tU,
expressed in either SI or U.S. customary
units, and for values of a from 5 to 858 at
58 intervals, can calculate (1) the normal
stress in the joint, (2) the shearing stress in
the joint, (3) the factor of safety relative to
failure in tension, (4) the factor of safety
relative to failure in shear, (5) the overall
factor of safety for the glued joint. (b) Apply
this pro
gram, using the dimensions and loading of
the members of Probs. 1.29 and 1.31,
knowing that sU 5 150 psi and tU 5 214 psi
P'
for the glue used in Prob. 1.29, and that sU
1.C5 A load P is supported as shown by
5 1.26 MPa and tU 5 1.50 MPa for the glue
two wooden members of uni form
rectangular cross section that are joined by used in Prob. 1.31. (c) Verify in each of
these two cases that the shearing stress
is maximum for a 5
Fig. P1.C5
458.
51 Computer Problems 1.C6 Member ABC is supported by a pin and bracket at A, and by two links that are pinconnected to the member at B and to a fixed support at
D. (a) Write a computer program to calculate the allowable load Pall for
any given values of (1) the diameter d1 of the pin at A, (2) the common
diameter d2 of the pins at B and D, (3) the ultimate normal stress sU in
each of the two links, (4) the ultimate shearing stress tU in each of the
three pins, (5) the desired overall factor of safety F.S. Your program should
also indicate which of the following three stresses is critical: the normal
stress in the links, the shearing stress in the pin at A, or the shearing stress
in the pins at B and D (b and c). Check your program by using the data
of Probs. 1.55 and 1.56, respectively, and comparing the answers obtained
for Pall with those given in the text. (d) Use your program to determine the
allowable load Pall, as well as which of the stresses is critical, when d1 5
d2 5 15 mm, sU 5 110 MPa for aluminum links, tU 5 100 MPa for steel
pins, and F.S. 5 3.2.
Top view
200 mm 180 mm
12 mm
8 mm
ABC
B
A
C
B
20 mm 8 mm
8 mm
P
DD
Front view
12 mm
Side view
Fig. P1.C6
This chapter is devoted to the study of
deformations occurring in structural
components subjected to axial loading.
The change in length of the diagonal
stays was carefully accounted for in the design of this cable-stayed bridge.
52
CHAPTER
Stress and Strain—Axial
Loading
53
and design of structures relates to the deforma
tions caused by the loads applied to a structure.
2.1 Introduction
Clearly, it is impor tant to avoid deformations so
2.2 Normal Strain Under Axial Loading
large that they may prevent the structure from
2.3 Stress-Strain Diagram
fulfilling the purpose for which it was intended. But
*2.4 True Stress and True Strain 2.5 Hooke’s Law;
the analysis of deformations may also help us in the
Modulus of Elasticity
determination of stresses. Indeed, it is not always
2.6 Elastic versus Plastic
possible to determine the forces in the mem bers of
Behavior of a Material
a structure by applying only the principles of statics.
2.7 Repeated Loadings;
This is because statics is based on the assumption
Fatigue 2.8 Deformations of
of undeformable, rigid structures. By considering
Members Under Axial Loading
engineering structures as deformable and analyzing
2.9 Statically Indeterminate
the deformations in their various members, it will be
Problems 2.10 Problems
pos sible for us to compute forces that are statically
Involving Temperature
indeterminate, i.e., indeterminate within the
Changes
framework of statics. Also, as we indicated in Sec.
2.11 Poisson’s Ratio
1.5, the distribution of stresses in a given member is
2.12 Multiaxial
Loading; Generalized statically indeterminate, even when the force in that
member is known. To determine the actual
Hooke’s Law
*2.13 Dilatation; Bulk distribution of stresses within a member, it is thus
necessary to analyze the deformations that take
Modulus 2.14
place in that member. In this chapter, you will
Shearing Strain
consider the deformations of a structural member
2.15 Further Discussions of
such as a rod, bar, or plate under axial loading.
Deformations Under Axial
First, the normal strain P in a member will be
Loading; Relation Among E, n,
defined as the deformation of the member per unit
and G
*2.16 Stress-Strain Relationships for Fiber-Reinforced length. Plotting the stress s versus the strain P as
the load applied to the member is increased will
Composite
yield a stress-strain diagram for the material used.
Materials
2.17 Stress and Strain Distribution Under Axial Loading; From such a diagram we can determine some
important properties of the mate rial, such as its
Saint
Venant’s Principle
modulus of elasticity, and whether the material is
ductile or brittle (Secs. 2.2 to 2.5). You will also see
2.18 Stress Concentrations
in Sec. 2.5 that, while the behavior of most
2.19 Plastic Deformations
materials is independent of the direction in which
*2.20 Residual Stresses
the load is applied, the response of fiber-reinforced
2.1 INTRODUCTION
com
In Chap. 1 we analyzed the stresses created in
posite materials depends upon the direction of the
various members and connections by the loads
load. From the stress-strain diagram, we can also
applied to a structure or machine. We also learned determine whether the strains in the specimen will
to design simple members and connections so that disappear after the load has been removed—in
they would not fail under specified loading
which case the material is said to behave
conditions. Another important aspect of the analysis elastically—or whether a permanent set or plastic
Chapter 2 Stress and Strain— Axial Loading
deformation will result (Sec. 2.6). Section 2.7 is
devoted to the phenomenon of fatigue, which
causes structural or machine components to fail
after a very large number of repeated loadings,
even though the stresses remain in the elastic
range.
The first part of the chapter ends with Sec. 2.8,
which is devoted to the determination of the
deformation of various types of members under
various conditions of axial loading.
In Secs. 2.9 and 2.10, statically indeterminate
problems will be considered, i.e., problems in which
the reactions and the inter nal forces cannot be
determined from statics alone. The equilib rium
equations derived from the free-body diagram of the
member under consideration must be
complemented by relations involving deformations;
these relations will be obtained from the geometry
of the problem.
54
In Secs. 2.11 to 2.15, additional constants associated with isotropic 55
2.2 Normal Strain under Axial Loading
materials—i.e., materials with mechanical characteristics independent
of direction—will be introduced. They include Poisson’s ratio, which
relates lateral and axial strain, the bulk modulus, which characterizes
the change in volume of a material under hydrostatic pressure, and
the modulus of rigidity, which relates the components of the shearing
stress and shearing strain. Stress-strain relationships for an isotropic
material under a multiaxial loading will also be derived.
In Sec. 2.16, stress-strain relationships involving several distinct
values of the modulus of elasticity, Poisson’s ratio, and the modulus
of rigidity, will be developed for fiber-reinforced composite materials
under a multiaxial loading. While these materials are not isotropic,
they usually display special properties, known as orthotropic proper
ties, which facilitate their study.
In the text material described so far, stresses are assumed uni
formly distributed in any given cross section; they are also assumed
to remain within the elastic range. The validity of the first assump
tion is discussed in Sec. 2.17, while stress concentrations near circu
lar holes and fillets in flat bars are considered in Sec. 2.18. Sections
2.19 and 2.20 are devoted to the discussion of stresses and deforma
tions in members made of a ductile material when the yield point of
the material is exceeded. As you will see,
Let us consider a rod BC, of length L and
permanent plastic deforma tions and
uniform cross-sectional area A, which is
residual stresses result from such loading
suspended from B (Fig. 2.1a). If we apply a
conditions.
load P to end C, the rod elongates (Fig.
2.1b). Plotting the magnitude P of
2.2 NORMAL STRAIN UNDER AXIAL B B L
LOADING
A
to the analysis of the rod
the load against the
under consider ation, it
P
deformation d (Greek letter
cannot be used directly to
C
delta), we obtain a certain
predict the deformation of a
load-deformation diagram
rod
(Fig. 2.2). While this diagram
con tains information useful C
same deformation in a rod B9C9 of the
same length L, but of cross-sectional area
2A (Fig. 2.3). We note that, in both cases,
of the same material but of different
dimensions. Indeed, we observe that, if a the value of the stress is the same: s 5 PyA.
On the other hand, a load P applied to a rod
deformation d is produced in rod BC by a
load P, a load 2P is required to cause the
B0C0, of the same
B' B' L
P
(a) (b)
Fig. 2.1 Deformation of axially-loaded rod.
Fig. 2.2 Loaddeformation diagram.
C'
Fig. 2.3
2P
C'
2A
56 Stress and Strain—Axial Loading cross-sectional area A, but of length 2L, causes a deformation 2d in that rod
(Fig. 2.4), i.e., a deformation twice as large as the deforma
tion d it produces in rod BC. But in both cases the ratio of the
B B 2L
d
P 5 L (2.1)
C
2
Plotting the stress s 5 PyA against the
deformation over the length of the rod is the strain P 5 dyL, we obtain a curve that is
characteristic of the properties of the
same; it is equal to dyL. This observation
brings us to introduce the concept of strain: material and does not depend upon the
We define the normal strain in a rod under dimensions of the particular specimen used.
This curve is called a stress-strain diagram
axial loading as the deformation per unit
length of that rod. Denoting the normal strainand will be dis
cussed in detail in Sec. 2.3.
by P (Greek letter epsilon), we write
Since the rod BC considered in the
preceding discussion had a uniform cross
section of area A, the normal stress s could
be
A
Thus, it was appropriate to
normal strain at point Q as
define the strain P as the ratio
of the total deformation d over
the total length L of the rod. In
Fig. 2.4
the case of a member of
variable cross-sectional area
A, however, the normal stress
s 5 PyA varies along the
member, and it is necessary to
define the strain at a given
point Q by considering a small
Q
element of undeformed length
C
Dx (Fig. 2.5). Denoting by Dd
d
¢x 5 d dx (2.2)
the deforma
P
assumed to have a constant tion of the element under the
value PyA throughout the rod. given loading, we define the
xx
¢d
P 5 lim ¢xy0
Fig. 2.5 Deformation of axially loaded member of variable cross
sectional area.
P
Q
x+ x +
Since deformation and length are expressed in
the same units, the normal strain P obtained by
dividing d by L (or dd by dx) is a dimensionless
quantity. Thus, the same numerical value is
obtained for the normal strain in a given
member, whether SI metric units or U.S.
customary units are used. Consider, for instance,
a bar of length L 5 0.600 m and uniform cross
section, which undergoes a deforma tion d 5 150
3 1026 m. The corresponding strain is
d
P 5 L 5 150 3 1026 m
0.600 m 5 250 3 1026 m/m 5 250 3 1026
Note that the deformation could have been
expressed in microme ters: d 5 150 mm. We
would then have written
d
P 5 L 5 150 mm
0.600 m 5 250 mm/m 5 250 m
and read the answer as “250 micros.” If U.S.
customary units are used, the length and
deformation of the same bar are, respectively,