CIV E 205 Mechanics of Solids II. Course Notes

Transcription

1 University of Waterloo Department of Civil Engineering CIV E 205 Mechanics of Solids II Instructor: Tarek Hegazi Room: CPH 2373 G, Ext tarek@uwaterloo.ca Course Web: Course Notes

2 CIV. E. 205 MECHANICS OF SOLIDS II COURSE OUTLINE Instructor: Dr. T. Hegazy Lectures: MWF 9:30 - CPH 3385 Office: CPH 2373G Ext.: tarek@uwaterloo.ca Web: T.A.s: University of Waterloo Civil Engineering Textbook: - Hibbeler, 2002 Mechanics of Materials, 5 th Edition, Prentice Hall. - Course Notes Download What is Covered: 1. - Introduction & Review 2. - Analysis of Stress 3. - Analysis of Strain 4. - Stress-Strain Relations 5. - Strain Energy 6. - Theories of Failure 7. - Deflection of Beams 8. - Energy Methods 9. - Principle of Virtual Work Influence Lines Reciprocal Theorem Buckling of Columns Special Beam Problems (Optional) Marking: 4 10% 40% Held on dates announced in class Final Examination: 60% Bridge Competition: Bonus Notes: - Each week, a number of suggested problems will be given to serve as background study for the quizzes. Solutions are not to be handed in. - Teaching Assistants will provide one-to-one help and will prepare you for quizzes. - Course notes, solutions to suggested problems, and solutions to quizzes will be posted on the course web site. Dr. Tarek Hegazy 1

3 Objectives: Mechanics of Materials - Solve Problems in a structured systematic manner; - Study the behavior of bodies that are considered deformable under different loading conditions; & - Analyze and design various machines / systems 1. Basic Concepts a) Resultants of Forces (i.e., one force equivalent to many) Concurrent Forces P1 Non-Concurrent Forces P1 P2 P3 P2 Calculating the Resultant Graphically: Force Polygon Method Note: R substitutes all its component forces. Draw forces subsequently. Each is drawn parallel to its line of action and length equal to value of force. Accordingly, resultant is an arrow that closes the polygon (value & direction measured graphically). Calculating the Resultant Analytically: The X and Y components of all forces are summed X = (x components of all forces); Y = (y components of all forces); Then, R = SQRT (X 2 Y 2 ); tan θ = Y / X b) Equilibrium of a system subjected to Forces (i.e., Resultant of all forces on the system = 0) Concurrent Forces Two Equilibrium Conditions: R (resultant) = 0 Non-Concurrent Forces Three Equilibrium Conditions: 1. X components of all forces = 0 1. X components of all forces = 0 2. components of all forces = 0 2. Y Y components of all forces = 0 3. M (moment at any point) = 0 Note: It is possible to put a system in equilibrium by calculating the Resultant of all forces acting on it then applying an opposite and equal force to that resultant. Dr. Tarek Hegazy 1

4 Example: can we set a system subject to P1 P1 R these two forces Step 1: get resultant (R) in equilibrium? Step 2: apply (- R) P2 P2 Example: Find the resultant and introduce a new force to set the system in Equilibrium P5 P4 P6 P5 P4 P3 P5 P3 P4 P6 P1 P2 P3 P2 P1 P6 P1 P2 Forces Resultant Equilibrium Example: The Polygon to the right is for a system of forces in Equilibrium. P6 - The resultant of forces P2, P3, P4, P5, P6, & P7 is - The resultant of forces P4, P5, P6, P7, P1, & P2 is P5 P4 P7 P3 P2 P1 Example: Find the forces acting on the cable. F1 F2 45 o 60 o = 45 o 60 o 6 tons Three force system Directions of top two Forces are assumed. 6 tons Dr. Tarek Hegazy 2

5 Solution: Graphically using force polygon or analytically using Equilibrium equations. F2 F1 6 tons X = 0 F2 COS 60 F1 COS 45 = 0 Y = 0 F2 SIN 60 F1 SIN 45-6 = 0 Solving both, F1 = t ; F2 = t Example: In the following system of parallel forces, determine the reactions Ra, Rb a b Ra Rb Analytically: Y components of all forces = 0 => Ra Rb 6 = 0 => Ra Rb = 6 M (moment at any point) = 0 => M a = Rb x = 0 or, Rb x 7.5 = 18.5 or Rb = 2.47 t Then, Ra 2.47 = 6 or Ra = 3.53 t c) Three-Forces Theorem Three non-parallel forces in Equilibrium must intersect in a common point. The force polygon of the triangle of forces shows a continuous loop. Example: Use the theorem of three forces to determine the reactions Ra a 5 m 4 tons b Rb 4 m Using the force polygon, the reactions are determined. Dr. Tarek Hegazy 3

6 Example: Use the theorem of three forces to determine the reactions on the beam. Polygon 1 for the resultant of Ra and Rc Polygon 2 for the equilibrium of the beam a 45 o 60 o 6 tons b Resultant of Ra and Rc Ra Rc 3 m 1 m 3 m Rb d) Types of Supports Supports exert reactions in the direction in which they restrain movement. Roller Support (restricts in one direction only and allows rotation) Rubber Hinged or Pinned Support (restricts in two ways and allows rotation) Fixed Support (restricts in two directions and also restricts rotation) Intermediate Pin or Hinge (Gives one extra condition) Gives extra condition M Right = 0 M Left = 0 Force in direction of member Examples: Dr. Tarek Hegazy 4

7 e) Structural Representation of Real Systems Dr. Tarek Hegazy 5

8 f) Stability & Determinacy of Structures - A stable structure can resist a general force immediately at the moment of applying the force. Unstable Does not return to original shape if load is released Stable - A statically determinate structure is when the reactions can be determined using equilibrium equations. 1. Beams: if r < c 3 Unstable r = unknown support reactions. if r = c 3 Statically determinate c = additional conditions if r > c 3 Statically Indeterminate r = 3 (two at hinge one at roller) c = 2 (two intermediate hinges), then, r < c 3 Unstable r = 5 (four at hinges one at roller) c = 2 (two intermediate hinges), then, r = c 3 Stable & Statically Determinate r = c = r =, then 2. Frames: j = No. of joints m = No. of membrs r = unknown support reactions c = special conditions r = 4 (three at fixed end one at roller) c = 0, then r > c 3 Stable & Statically Indeterminate if 3m r < 3j c if 3m r = 3j c if 3m r > 3j c Unstable Statically determinate Statically Indeterminate j = 3; m = 2; c = 0; r = 3 Then, Stable & Statically determinate j = ; m = ; c = ; r = Then, 3m r =, 3j 2 =, or 3. Trusses: j = No. of joints m = No. of membrs r = unknown support reactions if m r < 2j if m r = 2j if m r > 2j Unstable Statically determinate Statically Indeterminate j = ; m = ; r = Then, m r =, 2j =, or j = 8; m = 12; r = 3 Then, m 3 = 15 < 2j, or Unstable Dr. Tarek Hegazy 6

9 2. Analysis of Forces Forces and their effects at different points: Concentrated Load Distributed Load Rotation (Couple) P W t/m M L P1 Effect of a load on another point parallel to its axis = W x L L M = P1. L P P1 b L a Ma = 0 P P b P a Ma = 0 Effect of loads P b M = P x L a Ma = 0 Forces: V M T M P Shear = Force In the X-Section Plane Bending Moment = Couple Normal to Plane Torsion = Couple in the X-Section Plane Normal Force = Perpendicular to X-Section Example: Calculate reactions without applying equilibrium equations P P M = P.L /2 L/2 L/2 L/2 L/2 Dr. Tarek Hegazy 7

10 Example: Determine the forces at section BC 150 lb C M = 150x5 150 lb lb lb.in B 750 lb. in 150 lb 750 lb.in Example: Determine the forces at section A z 8 Note: When the structural system is:, then x A 10 the free end is a good starting point for the analysis. 800 lb 800 lb y T=800x14 Mx=800x10 14 A 500 lb 500 lb Mz=500x14 A lb T=800x14 Equilibrium equations for each segment: Mz=500x lb Mx=0, My=0, Mz=0 Fx=0, Fy=0, Fz=0 800 lb 14 Example: Determine the forces at section HK 500 lb Dr. Tarek Hegazy 8

11 Example: Determine the internal forces at section ABCD Example: Determine the internal forces at section C Example: Determine the forces acting at section D Example: Determine the internal forces on the section Dr. Tarek Hegazy 9

12 3. Internal Loadings on Beams & Frames Sign convention for internal forces (N.F., S.F., & B.M.) ive 3.1 Analytical Approach: Stability & Determinacy Reactions Axis Sections Signs N, V, & M Equations Draw Diagrams Equilibrium Conditions: X Components of all forces = 0 Y Components of all forces = 0 M At any point = 0 Extra condition at intermediate Pin: M r ight side only = 0 = M left side only Analysis of Shear & Moment Equations: F.B.D.: Between load changes, make a cut and put 3 (equal & opposite) internal forces on each side. P Segment 1 Segment 2 M M X X Xa Ya V Cut V Yb Apply Equilibrium Equations to this segment alone Apply Equilibrium Equations to this segment alone Example: Calculate the Reactions 10 t 4 t 8 t a 2 m 1 m 4 Ma 3 Xa Ya a 1 m 1 m 1 m 6 t Step 1: Cantilever beam is stable and statically determinate Step 2: Looking at the system as a whole (right figure) and applying Equilibrium equations, X Y Ma = 0, then Xa 6 = 0, or Xa = 6 t = 0, then Ya 4 8 = 0, or Ya = 12 t = 0, then Ma 4 x 1 8 x 3 = 0, or Ma = 28 m.t. Check OK Dr. Tarek Hegazy 10

13 Example: Calculate the Reactions 3 t 2 t / m Step 1: m = 5; r = 4; j = 6; c = 1 3m 4 = 19 = 3j 1 then, Frame is stable and statically determinate Step 2: Looking at the system as a whole and applying Equilibrium equations, Mb Mc = 0, then 3 x 11 2 x 8 x 7 2 x Xa 9 x Ya = 0 6 m left only = 0 = 3 x 4 Xa x 6 Ya x 2 2 t/m x 4 x 2 = 0 Solving both equations, we get Ya = t, and Xa = 0.76 t Xa a Ya c 2 m 2 m 4 m 3 m, then Yb b Xb 4 m 2 m Y X = 0, then Ya Yb = 0, or Yb = 2.72t = 0, then Xa - Xb = 0 then Xb = 0.76 t Check OK Example: Calculate the Reactions Mb Ma Y = 0 = =. = 0 = =. 2 m 4 m Ya = 0 =. = 0 or Ya = 36 - Yb 4 t 4 m 2 t / m Yb 12 m 4 m Solving these equations, we find Yb = 20 t ; Ya = 16 t ; Xa = Xb = 0 Check OK Example: Calculate the Reactions Mb Y = 0 = = 50x5-4 Ya =0, or Ya = 62.5 t = 0 = Ya Yb - 50 = 0 or Yb = t d 1 m 50 t c Mc right side only = 0 = 4 Yb - 3 Xb or Xb = - 50/3 t 3 m X = 0 = Xa - Xb = 0 or Xa = -50/3 t Check OK Xa 1 m 4 m Ya Yb Xb Dr. Tarek Hegazy 11

14 Examples: Calculate and draw the S.F.D. and the B.M.D. Reactions: Reactions: Sections: Sections: Solved examples 6-1 to 6-6 Note on simple beam with distributed load: Dr. Tarek Hegazy 12

15 3.2 Graphical Approach: Stability & Determinacy - Reactions N, V, & M Relations Draw Diagrams Examples on Page 12 Rules: 1- Shear curve is one degree above load curve 2- Moment curve is one degree above shear curve 3- Moment is maximum at point with shear = 0 4- Between any two points: (look at table) - Area under load = difference in shear - Area under shear = difference in moment - Slope of shear curve = - (load trend) - Slope of moment curve = shear trend Solved examples 6-7 to 6-13 Dr. Tarek Hegazy 13

16 Examples: Calculate and draw the S.F.D. and the B.M.D. ive a b From point a to point b: - Load curve = - Shear curve = - Moment curve = - Area under load = = difference in shear = - Area of shear = = difference in moment = - Shear at point of max. Moment = - Max. moment can be calculated from shear diagram = = - Slope of shear curve = - Slope of moment curve = From point a to point b: a b - Load curve = - Shear curve = - Moment curve = - Area under load = = difference in shear = - Area of shear = = difference in moment = - Shear at point of max. Moment = - Max. moment can be calculated from shear diagram = = - Slope of shear curve = - Slope of moment curve = Dr. Tarek Hegazy 14

17 ive Examples: For the problems in page 12, draw the N.F.D., S.F.D., & the B.M.D. 1 m d 50 t c 4 t 2 2 t / m 0 3 m /3 50/3 4 m 0 12 m 4 m 1 m 4 m Dr. Tarek Hegazy 15

18 Examples: Calculate and draw the S.F.D. and the B.M.D. ive Dr. Tarek Hegazy 16

19 4. Stresses due to Forces Internal Forces: Normal Moment Shear Torsion M V T P M Normal Force = Perpendicular to X-Section Bending Moment = Couple Normal to Plane Shear = Force In the X-Section Plane Torsion = Couple in the X-Section Plane Stresses - τ A Q = A. Y In narrow rectangular beams, τ max = 1.5 V / A Dr. Tarek Hegazy 17

20 Example: Draw the S.F.D. and the B.M.D. for the beam below and then determine the maximum normal stress due to bending. Example: Draw the S.F.D. and the B.M.D. for the beam below and then determine the maximum normal stress to the left and to the right of point D. The beam has a section modulus of 126 in3. Example: Page 8 of notes - Determine the internal stresses at points B & C 150 lb C 150 lb 750 lb.in B I = 4 x 10 3 / 12 ; σ B = / A 750 x 5 / I = 7.5 psi ; σ C = / A x 5 / I = psi Dr. Tarek Hegazy 18

21 Example: Calculate normal stresses at section d and also at the section just below c. First, we get the reactions. Example: Determine the internal stresses at points A, B, C, & D B C VQ / It = 0.5 MPa D A B V = 3 KN T = 3 KN My = 10.5 KN.m C B A C T.c / J = 15.3 MPa D D My = 10.5 lb.in M. x / I = 107 MPa C A B D A Dr. Tarek Hegazy 19

22 Example: Page 8 of notes = Example: The timber used in the beam below has an allowable stresses of 1800 psi (normal) and 120 psi (shear). Determine the minimum required depth d of the beam. From the S.F.D. and the B.M.D., maximum values of: Moment = 7.3 Kip.ft = 90 Kip.in; Shear = 3 Kips Design based on allowable Normal stress: Check Shear stress: Redesign based on allowable shear stress: >120 (unacceptable) Solved Problems 6-14 to 6-20, 7-1 to 7-3, 8-4 to 8-6 Dr. Tarek Hegazy 20

23 5. Transformation of Stresses - Member under tension only (P) in one direction, i.e., a normal stress. But, let s consider an inclined plane. σ θ = (P Cos θ) / (A / Cos θ) or Very important conclusions: - Under tension only, shear is automatically present at various planes. - The plane of maximum shear is when Sin 2θ = max or when θ = Maximum shear = σ x /2 = P / 2A - It is important to study stress transformation and shear failure. σ θ = σ x Cos 2 θ τ θ = ½ σ x Sin 2θ - Member under two dimensional stresses. Positive Signes y σ y τ xy σ x x Questions: Is this the maximum stress? If not, then What is the value of max. normal stress & its orientation? and What is the value of maximum shear stress & its orientation? y' σ y τ x y σ x x' θ x General Equations: 75 MPa Example: For the given state of stress, determine the normal and shearing stresses after an element has been rotated 40 degrees counter-clockwise. 60 MPa 30 MPa σ x = 30 MPa ; σ y = -75 MPa ; τ xy = 60 MPa ; θ = 40 Applying the above equations, we get: σ x = 45.7 MPa ; σ y = MPa ; τ x y = MPa y' x' 40 x Solved Examples 9-2 to 9-6 Dr. Tarek Hegazy 21

24 Important Observations: 1. σ x σ y = σ x σ y = Constant Sum of normal stress is constant (90 degrees apart) for any orientation. 2. The plane in which shear stress τ x y = 0 is when: or tan 2θ = 2 τ xy / (σ x - σ y ) or at θ 1, θ 2 having 90 degrees apart. These are called principal planes. = 0 3. σ x becomes maximum when dσ x / dθ = 0, or when differentiating the following equation: we get, tan 2 θp = 2 τ xy / (σ x - σ y ) or, exactly at the principal planes, which has shear stress = 0. The value of the principal normal stresses are: σ max, min = σ x σ y ± 2 ( σ x - σ y ) 2 τ 2 xy 2 4. Since σ x σ y = constant, then, at the principal planes, σ x is maximum but σ y is minimum. 5. τ x y is maximum when planes, dτ / dθ = 0, or when: tan 2 θs = - (σ x - σ y ) / 2 τ xy and the value of maximum shear stress τ xy is: τ x y max = ( σ x - σ y ) 2 τ 2 xy 2 6. Similar to single stress situation, maximum is when dτ / dθ = 0, or when: θ =. Example: Check rule 1 for the example in previous page. In the general equations, even if the original τ xy on the element = 0, then still the shear at any plane (τ x y )has a value as a function of normal stresses. Dr. Tarek Hegazy 22

25 Example: Determine the maximum normal and shear stresses at point H. Forces at the section: Stresses at Point H: Principal stresses: Example: Determine the maximum normal and shear stresses at points H & K. Dr. Tarek Hegazy 23

26 Example: Determine the maximum normal and shear stresses at points H & K. Dr. Tarek Hegazy 24

27 Circular representation of plane stresses (Mohr s Circle): y Step 1: Given a state of stress, with σ x and σ y having 90 degrees apart. σ y τ xy σ x Step 2: Let s plot the σ x and σ y on a horizontal line Notice, the 90 degrees are now 180 apart. x 0 σ y θ σ x σ Step 3: Let s plot the σ x and σ y on a horizontal line then τ xy vertically at points 1 and 2 using signs. 0 τ Y (?,? )? τ xy σ y???? σ x τ xy X (?,? ) σ σ y Y X τ xy σ x Step 4: Draw a circle from the center to pass by points 1 and 2. Determine σ max, σ min, θ p, τ max, θ s? σ y Y Y X τ xy σ x 0? σ y σ x?? σ X Solved Examples 9-7 to 9-13 τ Dr. Tarek Hegazy 25

28 Example: For the given state of stress, determine the normal and shearing stresses after an element has been rotated 40 degrees counter-clockwise. 75 MPa 60 MPa 30 MPa Y (-75,60) X θ2 θ1 80 o 30 σ Y X Y X (30, -60) From the figure: Average stress = Center of circle = (30 75)/2 = -22.5, R = sqrt ( ) = 79.7 tan θ1 = 60 / 52.5, then θ1 = 48.8 o and θ2 = 80 - θ1 = 31.2 o Then, points X and Y have the following coordinates: σ x = R cos θ2 = * = 45.7 MPa σ y = R cos θ2 = MPa ; τ x y = R sin θ2 = MPa Principal stress values: σ max, σ min = Average ± R = ± 79.7 = 57.2, Example: For the given state of stress, determine: a) principal planes; and b) principal stresses. Analytically: σ x = -40 MPa; σ y = 60 MPa ; τ xy = 25 MPa tan 2θ p = 2 τ xy / (σ x - σ y ) = 2 x 25 / (-40-60) = -0.5 or at θ p1 = ; θ p2 = 76.7 σ max, σ min = Average ± R = Graphically: Two points X & Y Center = R = σ max, σ min = Average ± R Y σ x σ y ± 2 X (-40, -25) ( σ x - σ y ) 2 τ 2 xy 2 2θ = 10 ± 55.9 MPa Y (60,25) σ 60 MPa 25 MPa 40 MPa Y X 40 x Dr. Tarek Hegazy 26

29 Example: You have a Mohr circle of stress as shown below for two separate points. Draw the stresses on each element and its orientation from principal planes. 40 A (50,? ) R =? σ Y A (50, y ) 40 y 50 x = 50 Y 16.8 R = 60 Y = sqrt ( ) = 33.2 Tan 2θ = y / 50 or 2θ = 33.6 o θ 160 σ A (150, -y ) Y 100-X X 40 B (100-X, 40 ) y = 40 R = 60 x 100 B (?, -40 ) 2θ 160 B (100 X, -40 ) X = sqrt ( ) = 44.7 Tan 2θ = 40 / X or 2θ = 83.6 o σ Special Cases: 1. Case of pure tension A member under one directional stress σ x = P / A and σ y = 0 Let s use Mohr s circle. From Mohr s circle, maximum shear τ max = σ x / 2 = P/ 2A Maximum shear is at 45 degrees. τ max Y X 0 σ x y' σ y Y X τ max σ x σ x x' Case of pure torsion A member under only torsional stress, with σ x = 0 and σ y = 0, Let s draw Mohr s circle. Y X τ xy Y As seen, torsional stress creates normal Stresses which are maximum at 45 degrees. 0 σ max σ X Dr. Tarek Hegazy 27

30 Graphical representation of principal planes: Properties of a circle: an angle 2θ at the center of the circle, corresponds to an angle θ at the circumference. θ θ = 90 2θ 2θ = 180 Special case 5 Y (5,20) 20 Y X θ p 30 σ X (30, -20) Plane of principal normal stress Plane of maximum shear stress Y σ = -6.1 min Y 5 2θ p 30 5 σ max = θ s 30 θ p X θ s X Dr. Tarek Hegazy 28

31 3-Dimensional stress systems: (Absolute maximum shear stress) Assume σ1 > σ2 > σ3 are principal normal stresses ( no shear), then let s draw Mohr s circle. τ max σ 2 σ 1 σ3 σ2 σ1 σ σ 3 Note : Even if σ3 = 0, 3-D stress analysis becomes essential. Case 1: both σ1 and σ2 are positive Case 2: both σ1 and σ2 are negative Then, τ max = σ1 / 2 Then, τ max = σ2 / 2 τ max τ max σ2 σ1 σ2 σ1 Case 3: σ1 and σ2 have opposite signs Then, τ max = (σ1 - σ1) / 2 τ max σ2 σ1 Examples Dr. Tarek Hegazy 29

32 6. Transformation of Plain Strain - A structure should be designed so that its material and cross sectional dimensions can resist the maximum normal and shear stresses imposed on it. Equally important also that the structure does not deform much under the load, i.e., the ability to resist strains is crucial to the serviceability of structures. - Normal Strain (due to axial load bending moment) and Shear Strain (due to transverse shear torsion). Normal Strain Shear Strain = Questions: Strain = ε = Unitless = L / L Positive Signs (elongation and angle) Is this the maximum strain? If not, then What is the value of maximum normal strain and the plane in which it exists? and What is the value of maximum shear strain and the plane in which it exists? - General equations for strains on a plane at angle θ for a member under two dimensional strain. Notice that all equations look the same as those of stress transformation, except that τ xy is resembled by : General Equations: Given the three constants,, then, Normal strain at any angle θ: Shear strain at any angle θ: Principal (Normal) Strain: Orientation: Max. Value: Shear strain at this plane: Zero Maximum Shear Strain: Orientation: Max. Value: Solved Problems 10-1 to 10-8 Normal strain at this plane: Dr. Tarek Hegazy 30

33 - Strains before and after transformation: Positive Strains Negative Strains Positive Strains at positive angle θ at negative angle θ at θ = 0 θ Important Observations: 1. ε x ε y = ε x ε y = Constant (90 degrees apart) for any orientation. 2. The plane in which shear strain γ x y / 2 = 0 is when: = 0 or tan 2θ = γ xy / (ε x - ε y ) or at θ 1, θ 2 having 90 degrees apart. These are called principal planes. 3. ε x becomes maximum when dε x / dθ = 0, or when differentiating the following equation: we get, tan 2 θp = γ xy / (ε x - ε y ) or, exactly at the principal planes, which has shear strain = γ x y is maximum when dγ / dθ = 0, or when: tan 2 θs = - (ε x - ε y ) / γ xy 5. Similar to single stress situation θs = 45 o from θp. 6. Mohr s circle of strain: γ/2 Y (εy, γ/2) Y X 2θ p ε X (εx, - γ/2) ε min?, ε max?, γ max?, θ p?, θ s? Dr. Tarek Hegazy 31

34 Example: Given ε x = -200 x10-6, ε y = 1000 x10-6, γ xy = 900 x10-6. Find the strains associated with x y axes inclined at 30 degrees clockwise. Find principal strains and the maximum shear strain along with the orientation of elements. Solution First, we sketch the element with the given strains, as follows. Y X Then, we define two points X and Y to draw Mohr s circle. γ/2 x10-6 Shorter in X Longer in Y ive shear strain X (-200, -450) R= Y (1000, 450) 600 ε x10-6 2θ p R = Sqrt (60^2 450^2) = 750 Principal Strains: εmax, εmin = 400 ± 750 = 1150 x10-6, -350 x10-6 γx y at principal planes = 0 2θ p = tan -1 (450 / 600) = 36.8 o Max Shear Strains: γmax / 2 = R = 750 x10-6 εx = εy at Max shear plane = 400 x10-6 2θ s = 36.8 o 90 = o γ/2 x10-6 α = = 23.2 Then εx = R Cos α = x Cos 23.2 = -290 x10-6 γx / 2 = R Sin α = 750 Sin 23.2 = 295 x10-6 X (, ) R=750 α ε x o α 400 Y εy = 400 R Cos α = x Cos 23.2 = 1090 x10-6 γy / 2 = -R Sin α = -750 Sin 23.2 = -295 x10-6 X R=750 Y (, ) At 30 o Clockwise At Principal Planes At Maximum Shear Plane Y Y X θ=30 o Shorter in X Longer in Y -ive shear strain (clockwise rotation) θp=18.4 o Shorter in X Longer in Y No Shear strain θs= o X Longer in X Longer in Y -ive shear strain Dr. Tarek Hegazy 32

35 Absolute maximum shear strain Assume ε1 > ε2 > ε3 are principal normal strains (no shear), then let s draw Mohr s circle. γ/2 γ max /2 Note: Even if ε3 = 0, 3-D analysis is essential. Case 1: both ε1 and ε2 are positive Then, γ max /2 = ε1 / 2 ε3 ε2 ε1 ε γ/2 γ max /2 ε2 ε1 γ max /2 = (ε max - ε min ) / 2 Case 2: both ε1 and ε2 are negative Case 3: ε1 and ε2 have opposite signs Then, γ max /2 = ε2 / 2 Then, γ max /2 = (ε1 - ε2) / 2 γ max /2 γ/2 γ/2 γ max /2 ε2 ε1 ε2 ε1 Solved Examples: 10-1 to 10-7 Dr. Tarek Hegazy 33

36 Strain Measurements Using Strain Rosettes: - 45 o strain rosette versus 60 o strain rosette - Cemented on surface - Its electrical resistance changes when wires are stretched or compressed with the material being studied - Resistance changes are measured and interpreted as changes in deformation - Three values to get the state of strain at the point - Automated condition assessment of bridges - Check the strains on older structures or εx = εx cos 2 θ εy Sin 2 θ γxy cos θ. Sin θ Readings: εa, εb, εc At: θa=0, θb=45, θc=90 Unknowns: εx, εy, γxy Applying into the general equation: εx = εa εy = εc γxy= 2εb (εa εc) Readings: εa, εb, εc At: θa=0, θb=60, θc=120 Unknowns: εx, εy, γxy Applying into the general equation: εx = εa εy = (2εb 2εc - εa) / 3 γxy= 2(εb εc) / Sqrt(3) Substitute into either equation 3 times using εa, εb, εc to get the unknowns εx, εy, γxy at the measurement point. Example: Using the strain rosette shown, the measured values at each stain gauge is as follows: b a εa = 8 x 10-4, εb = -6 x 10-4, εc = -4 x 10-4 Determine the principal strains at the point. c Solution Using Equations: θa = 90, θb = 135, θc = 180 Applying into the general strain transformation equation: εa = 8 x 10-4 = εx cos 2 90 εy Sin 2 90 γxy cos 90. Sin 90 εb = -6 x 10-4 = εx cos εy Sin γxy cos 135. Sin 135 εc = -4 x 10-4 = εx cos εy Sin γxy cos 180. Sin 180 Then: εy = εa = 8 x 10-4 ; εx = εc = -4 x 10-4 ; γxy /2 = 16 x 10-4 Using Mohr s circle, we determine principal strains: ε1 = 12 x 10-4 ; ε2 = -8 x 10-4 γ/2 x10-4 A R=10 16/2-4 ε x /2 C Solution Using Only Mohr s Circle: γ/2 x10-4 A Directions (a) and (c) are 90 degrees apart This means that the center of the circle is the Average strain = (εa εc) / 2 = (8 x x 10-4 ) /2 = 2 x 10-4 From the two triangles shown, d =, Then, R = Sqrt( d ) = As such, B -6 2θp R R 8 2θp 2 6 d =? 8 ε x10-4 ε1 = 2 R = 12 x 10-4 ; ε2 = 2 R = -8 x 10-4 C Dr. Tarek Hegazy 34

37 Example: Solution: Strategy: We draw a Mohr s circle for strain and on it will find the strains at the orientations of the strain gauges (45o apart). Y X? Longer in X Shorter in Y ive shear strain.?? Dr. Tarek Hegazy 35

38 7. Relationship between Stress & Strain - A stress in one direction causes elongation in its direction and shortening in the other two depending on the material s Poisson s ratio (ν). Generalized Hooke s law εx = σx / E εx = - ν. σy / E εx = εy = - ν. σx / E εy = σy / E εy = εz = - ν. σx / E εz = - ν. σy / E εz = - Assumptions: (1) τ has not correlation with εx and εy; (2) σx and σy have no relation with γxy ; (3) principal strains occur in directions parallel to principal stresses. -General Equations: E.εx = σx - ν (σy σz); E.εy = σy - ν (σx σz); E.εz = σz - ν (σx σy); G. γxy = τxy G. γyz = τyz G. γzx = τzx - Relationship between E, ν, G: Let s consider the case of pure torsion, i.e., σ x = 0 and σ y = 0, Let s draw Mohr s circles for both stress and strains. τ Y γxy/2 Y 0 σ max σ 0 ε max ε X X Principal stresses are: σ 1 = τxy ; σ 2 = - τxy Principal strains are: ε 1 = γxy/2 ; ε 2 = - γxy/2 Now, let s apply Hook s Equation, as follows: E.ε1 = σ1 - ν (σ2) ; then E.ε1 = E. γxy/2 = τxy - ν (- τxy) = τxy. (1 ν) = G. γxy. (1 ν) Then, Bulk Modulus G = E / 2 (1 ν) K = E / 3 (1-2 ν) Note: Since most engineering materials has ν = 1/3, then G = 3/8 E and K = E Dr. Tarek Hegazy 36

39 Example: Notice the difference between Mohr s circles for stress & strain Example: Example: Example: Dr. Tarek Hegazy 37

40 Example: Example: Match each one of the following Mohr s circle for stress with a Mohr s circle for strain. Explain. Example: A 60 degree strain rosette is put at point A. The strain guage readings are: εa = 60x10-6 εb = 135x10-6 εc = 264x10-6 a) Determine the Principal strains at point A and their directions. b) If the bracket is made of steel for which E = 200 GPa and ν = 0.3, determine the Principal stresses at point A. Solution: Approach: from strain rosette readings, we get strains at the point, then calculate principal strains and finally convert these principal strains into principal stresses. Dr. Tarek Hegazy 38

41 a) First, we determine the normal and shear strains at point A from rosette readings. Accordingly, Second, we determine principal strains using Mohr s circle. Center of circle = average normal strain = (60 246) / 2 = 153x10-6 Accordingly, Principal strains = 153 ± R = 272 x10-6 ; 33 x10-6 b) Using Hooke s law with σ 3 = 0 Another Solution: Approach: from strain rosette readings, we get strains at the point, then we convert them into stresses and finally calculate principal stresses. First, we calculate strains at point A, same as above. Then, Now, we convert these strains to stresses: Dr. Tarek Hegazy 39

42 We now can use Mohr s circle for stress to determine Principal stresses: Example Solution Approach: Since we are given the forces, let s calculate the Stresses at point P, then, convert these stresses into strains. Forces on Section at P. Forces at end of beam. Stresses at Point P: Normal stresses Dr. Tarek Hegazy 40

43 Shear Stresses Strains at Point P: Example: Forces at section of Point A: Stresses at point A: ` Strains at point A: Dr. Tarek Hegazy 41

44 Mohr s circle of Strain: Strain Gauge Readings: Solved Problems 10-9 to Dr. Tarek Hegazy 42

45 8. Theories of Failure Ductile Material (Yield Failure) All theories deal with PRINCIPAL STRESSES Brittle Material (Fracture Failure) - Max. normal stress (Rankin s Theory) - Max. shear stress (Tresca Criterion) - Max. Energy of Distortion (Von Mises Criterion) - Other: Max. principal strain (St. Venant) Max. normal stress (Rankin s Theory) Principal stresses σ2 σ1 Failure when: σ1 > σy / F.S. or σ2 > σy / F.S. where F.S. > 1 σ2 σy A state of plane stress Is safe inside the square And unsafe outside (assume tension = compression) -σy -σy σy σ1 Max. shear stress (Tresca Criterion) A specimen under tension reached maximum stress σy, then, the maximum shear that the material can resist is σy /2 from Mohr s Circle. Then, failure is when τ Absolute max. shear (3-D analysis) = σ max σ min /2 > σy / (2 * F.S.) Dr. Tarek Hegazy 43

46 Energy of Distortion (Von Mises Criterion) To be safe, U d on element < U d yield U = ½ σ.ε For the 3-D stress Case: 1 [(σ1 σ2) 2 (σ2 σ3) 2 (σ3 σ1) 2 ] < 2 σ 2 yield 12G 12G or Simply, (σ1 σ2) 2 (σ2 σ3) 2 (σ3 σ1) 2 < 2 σ 2 yield For the 2-D stress Case: (σ3 = 0) (σ1 2 σ1 σ2 σ2 2 ) < σ 2 yield Other: Max. principal strain (St. Venant) Rarely used Using Hooke s law E ε max = σ1 ν (σ2 σ3) < σ yield Fracture of Brittle Materials Brittle materials are relatively weak in Tension. Failure criterion is Maximum Principal Tensile Stress. Under Tensile force, failure is due to tension. Under Torsion, failure is still due to tension at an angle. Element is safe when: Example: Twist of a piece of chalk. Solved Examples to Dr. Tarek Hegazy 44

47 Example: A steel shaft (45 mm in diameter) is exposed to a tensile yield strength = σyield = 250 MPa. Determine P at which yield occurs using Von Mises and Tresca critera. Solution 1) Principal Stresses σx = P / A = P /π (0.0225) 2 τxy = T.c / J = 1.7 x (0.0225) / ½ π (0.0225) 4 = Mohr's circle: Center = σx / 2 R = [(σx/2) 2 τxy 2 ] ½ σ1 = σx / 2 R; σ2 = σx / 2 - R T = 1.7 KN.m P? σx 2) Using Von Mises σ1 2 σ1σ2 σ2 2 = σ yield 2 (σx / 2 R) 2 (σx / 2 R) (σx / 2 - R) (σx / 2 - R) 2 = σ yield 2 (σx/2) 2 3 R 2 = σ yield 2 σx 2 3 τxy 2 2 = σ yield, substituting with R,, substituting with σx & τxy, X [P /π (0.0225) 2 ] 2 3 x (95.01) 2 2 = σ yield = then, P = KN 3) Using TRESCA σ1 and σ2 have opposite signs, then τ max (3-D) = σ1 - σ2 / 2, which reaches failure of τ yield = σ yield / 2 (σx/2 R) - (σx/2 - R) / 2 = σ yield / 2 then, R = σ yield / 2, substituting with R and squaring both sides, (σx/2) 2 τxy 2 = (σ yield / 2) 2, substituting with σx & τxy, [P /2 π (0.0225) 2 ] 2 (95.01) 2 = then, P = KN Notice the force P under TRESCA (focuses on Shear) is smaller than Von Mises Dr. Tarek Hegazy 45

48 Deflection of Beams and Shafts Beams and shafts deflect under load. For serviceability, we need to make sure deflection is within allowable values. Also, the shape of the beam under the load (elastic curve) needs to be studied. Terminology: - EI = Flexture rigidity or Bending Stiffness - R = Radius of Curvature - 1/R = Curvature υ - Hookes Law: 1/R = M / EI - The elastic curve: dθ R y R Rdθ = ds dx or 1/R = dθ/dx Also, dυ/dx = tan θ θ Differentiating both sides, then d 2 υ/dx 2 = dθ/dx θ ds θdθ ds θ dx dυ Notes: Accordingly, 1 = M = dθ = d 2 υ R EI dx dx 2 - Integration of (M/EI) determines the slope of the elastic curve: - Double integration of M/EI determines the deflection: - Recall relationships between load, shear, and bending moment. Now, we can expand it to: EI d 2 υ/dx 2 = M(x); EI d 3 υ/dx 3 = V(x); EI d 4 υ/dx 4 = -W(x) Determining the integration constants C1 and C2: Substituting at points of known deflection and/or slope, we can determine the constants of integration. Dr. Tarek Hegazy 46

49 Shape of Elastic Curve: inflection point at location where moment=0 Calculating Slope & Displacement by Integration: Step-by-Step 1. Get beam reactions: X = 0, Y = 0, M = 0 2. Get equation of B.M. at each beam segment with change in load or shape 3. Integrate the moment once to get the slope 4. Integrate the moment a second time to get the deflection (elastic curve) 5. Substitute at points of special conditions (boundary conditions) to get the constants C1 & C2 6. Rewrite the slope and deflection equations using the constants 7. Put slope = 0 to determine the location (x) that has maximum deflection Example: For the part AB, determine the equation of the elastic curve and maximum deflection if: I = 301x10 6 mm 4, E=200 GPa, P=250 KN, a = 1.2 m, L = 5 m. P Ya L Yb a Dr. Tarek Hegazy 47

50 Solution 1. Reactions: M a = 0 Yb. L - P. (L a) = 0 or Yb = P (1 a/l) P Y = 0, then Ya Yb P = 0 or Ya = - P. a/l P.a/L P(1 a/l) 2. Bending moment equations: Mx x 0 to L V P.a/L Mx = - P.a.x / L Fx Fx P Mx x = 0 to a V Mx = - P. x 3. Integrate the moment to get the slope: 8. Applying same steps at the free end: = - P.a.x 2 / 2L C1 = - P.x 2 /2 C3 (3) 4. Integrate a second time to get the (elastic curve) = - P.a.x 3 / 6L C1. x C2 = - P.x 3 / 6 C3. x C4.(4) 5. Substitute at points of known conditions 6. Final equations: at support A: [x = 0. y = 0] then, C2 = 0 also, at support B: [x = L, y = 0] then, 0 = - P a L 3 / 6L C1. L or C1 = P.a.L/6 = - P.a.x 2 / 2L P.a.L /6 (1) = - P.a.x 3 / 6L P.a.L.x /6... (2) 7. Put slope = 0 at maximum deflection 0 = - P.a.x 2 / 2L P.a.L /6 get x = 0.577L Slope at B right = Slope at B Left Slope left = using equation (1), x=l = -P.a.L/2 P.a.L/6 Slope right = using eq. (3), x=a = -P.a 2 /2 C3 we get C3 Also, at B: [x = a, y = 0] Using Equ. (4), we get C4 Using this value in equation (2), we get Max deflection = 8 mm Up. Solved problems 12-1 to 12-4 Dr. Tarek Hegazy 48

51 Calculating Slope & Displacement by Moment Area Method: 1 st Moment Area Theorem: Recall M = dθ EI dx Then, θ B/A = X A X B change in slope M EI dx area under M/EI diagram 2 nd Moment Area Theorem: t BA = (vertical distance from B on elastic curve to tangent at A) = Moment of the area under M/EI around point B. d B = d B. M dx EI X A X B Note: t AB = t BA Case 1: Cantilever Notice that tangent at point A is horizontal. -Deflection at any point: -Slope at any point: θ B/A = M dx = θ B - θ A = θ B EI X A X B θ A = Case 2: Symmetric Loading Option 1 Deflection is max at mid beam (C). At this point θ C = -Deflection at any point: -Slope at any point: X d θ B/A = M dx = θ D - θ C = θ D EI X C A C θ C = C = d d t dc B t BC Dr. Tarek Hegazy 49

52 Case 3: Unsymmetrical Loading Option 2 L1 L2 -Deflection at any point: d t DA = t BA. L1/(L1L2) A d B -Slope at any point to the right: d θ D/A = M dx = θ D - θ A with θ A being negative = θ D θ A EI X a X d t BA / (L1L2) A t DA t BA A L1 d θ A d L2 B θ D θ A t DA t BA A -Slope at any point to the left: θ A θ D/A = M dx = θ D - θ A, both negative = θ A - θ D EI X a X d t BA / (L1L2) θ D Case 4: Over-Hanging Beam L1 θ B = t AB / L1 A heavy load θ B B C t CB c = ( c t CB ) / L2 L2 t AB Then, c = θ B. L2 - t CB t BA t CA = t BA. (L1L2)/L1 A θ A B C Then, c = t CA - L1 L2 heavy load c Dr. Tarek Hegazy 50

53 Case 5: Unsymmetrical Loading Point of Max. Deflection θ B = t AB / L A x =? B t BC X B θ B/C = M dx = θ B - θ C = θ B = t AB / L EI X C We get x, then C = max = t BC t BA C θ C = 0 C = max Note: Equivalence in Bending Moment Diagrams a -b a = = a ive -ive -b -a Method of Superposition: - Using Standard tables for various beam conditions and types of loads (Appendix C) - Adding up deflections caused by individual loads Solved Problems: 12-7 to Dr. Tarek Hegazy 51

54 Example: Determine θ C and A A 150lb 4 L1 300lb C D B lb 400 lb 600 D M/EI 600 t DC =1/EI [ (600 x 24 /2). 2/ (1800 x 24 /2). 1/3. 24 ] θ C = t DC / 24 = ( A t AC ) / 4 A θ C t DC A 600 t AC = 1/EI [600. 4/2. 2/3. 4] t AC Example: Determine θ A and D A B wl/6 D wl/3 θ A = t BA / L = Moment of B / L B.M.D. 3 rd degree = [w.l 2 /6EI. L/2. L/3 - w. L 2 /6EI. L/4. L/5] / L = 7 w.l 3 / 360EI wl 2 /12 = wl 2 /6 Also, d t DA = t BA / 2 Then A θ A 3 rd degree d d L/5 -wl 2 /6 B d = t BA / 2 - t DA = 7 w.l 3 /720EI L/2 -wl 2 /48 - [1/2.w.L 2 /12EI.L/2.L/6 - ¼. wl 2 /48EI. L/10] = 5 w.l 4 / 768EI t DA t BA Dr. Tarek Hegazy 52

55 Using Deflection Calculations to Solve Statically Indeterminate Beams P P = 1 2 =0 R B? statically indeterminate P = M A? θ1 θ2 =0 First, we reduce the beam to a statically determinate, then We compensate for the change in the deflection behavior. Example: A 6 ft 12 Kips 3 Kip/ft B 6 ft 12 ft Determine the reactions, then draw the S.F.D. & the B.M.D. C 12 Kips = 3 Kip/ft 1 A 1 B C t AC t BC R B 2 2 = - R B L 3 / 48 EI 1 2 =0 Dr. Tarek Hegazy 53

56 Example: Solution: Dr. Tarek Hegazy 54

57 Energy Methods - For a structural element under load and deformation, External Work U e = Internal Strain Energy U i. - External work U e is a function of the load P and deflection. (deflection is at same point and direction of load) U e = ½ P. - Also, the Internal strain energy in the structure U i is a function of the stress σ and strain ε in the element, summed over the volume of the structure. U i = ½ σ. ε. V = σ 2 Strain Energy per unit volume 2E. V σ ε Normal Shear L U i = σ 2. dv and U i = τ 2. dv V = dv = da dx 2 E 2 G v v v A 0 L when A is constant, Observe the units. Strain Energy calculations for different loading conditions are shown in next page. V = A dx 0 Determining Deflections Using Conservation of Energy Single External load Deflection in the direction of load: U e = U i U e = ½ P. & U i = U e = ½ M o. θ & U i U e = ½ P. & U i Limitations: Applies to single load only. Also, in case 2, only solpe is calculated not deflection. Also, how to get deflection at a point at which no direct load is applied. Solved Examples 14-1 to 14-7 Dr. Tarek Hegazy 55

58 Strain Energy Calculations Axial Load U i = σ 2. dv = N 2 da dx 2E 2EA 2 v A 0 L Normal Stress Example: Truss with varying axial loads on individual members. (Cross section area A is constant, then V = A. L) Bending Moment L U i = σ 2 /2E dv or U i = M 2. y 2 da dx v 2 E I 2 A 0 σ = M.y I = I L U i = M 2 dx 2 E I 0 Shear Stress Pure Shear U i = ½ τ. γ. dv = τ 2 /2G A. dx v 0 τ = V.Q L I.t L U i = fs V 2 dx where, fs = 6/5 - rectangular section 2 G A 0 Torsion U i = ½ τ. γ dv = τ 2 /2G. da. dx v A 0 L τ = T.c J U i = T 2 dx 0 2 G J L Dr. Tarek Hegazy 56

59 Example: Determine the strain energy due to both shear and bending moment in the following cantilever. The cross section is a square of length a, with EI being constant. w P Example: Determine deflection at C, neglect shear strain energy.. 2EI EI L/2 L/2 C Dr. Tarek Hegazy 57

60 Principle of Virtual Work Conservation of Virtual Work Work-Energy method is not able to determine the deflection at a point at which no direct load exists.? Solution: Put a virtual load of 1.0 at the desired point of a virtual system. Then apply the principal of conservation of virtual work, as follows: Real Beam Virtual Beam 1.0? External Virual WorK = Internal Virtual Energy ½ Virtual load x Real displacement = ½ Virtual Stress x Real Strain x Volume L 1.0 x = σ V. ε R. V = σ V. σ R. da dx E A 0 L L = n N A dx = n N dx Axial A A E 0 0 A E Load L m M dx Bending 0 E I L fs v V dx Shear 0 GA L t T dx Torsion 0 G J Examples: Real Virtual Determine slope at desired point 2. Determine horizontal deflection at desired point Solved Examples to Real N M V T Virtual n m v t 1.0 Dr. Tarek Hegazy 58

61 Example: Determine the deflection at mid span. w Example: Get deflection at A. w L/2 L/2 A Dr. Tarek Hegazy 59

62 Example on Virtual Work - determine the vertical deflection at point A. A 3 t/m Determine the horizontal deflection at point A. A Dr. Tarek Hegazy 60

63 Calculate: The horizontal displacement at point B. Vertical displacement at point C. Slopes at points A, C, and D. Relative displacement between points E and F. EI = Constant 3 m 4 KN 2 KN / m D E 4 KN / m C F 4 m A B 4 m 2 m 4 m 2 m Dr. Tarek Hegazy 61

64 Suggested Problems Calculate the vertical deflection at point d. EI = 20,000 m2.t Calculate: - The horizontal displacement at point b, - The vertical displacement at point g - The slope at point f EI = 20,000 m2.t For the shown frame calculate: - The vertical deflection at point c - The horizontal deflection at point d EI = 15,000 m2.t For the second problem, assume support B is hinged. In this case, draw the S.F.D. and the B.M.D. for the frame. Dr. Tarek Hegazy 62

65 Important Use of the Virtual Work method: SOLVING statically indeterminate structures A w B C L/2 L/2 1 R B 2 = 0 Reduced System Compensation w = 1 2 R B. 1.0 Also, w L/2 L/2 = Reduced System θ 1 w θ 1 M A θ 2 = 0 Compensation M A. 1.0 θ 2 Examples: Dr. Tarek Hegazy 63

66 Calculating Deflections Using Castigliano s Theorem - Put an external load at the position of required deflection: external load (Q) either horizontal or vertical to get horizontal or vertical deflection; or an external moment to get slope. - Deformation = first derivative of the Strain Energy with respect to the applied load. = du / dq, & substituting Q = 0 = δ N 2 dx = N δ N dx Axial Load (Trusses) δq 2EA EA δq = δ M 2 dx = M δ M dx Bending Moment δq 2 E I E I δq 0 = δ fs V 2 dx = fs V δ V dx Shear δq 2 G A G A δq 0 0 = δ T 2 dx = T δ T dx Torsion δq 2 G J G J δq 0 0 L L L L L L L L Example: Determine the horizontal deflection at point B. Cross-section area= 12 in2 E= psi. AB = 48 in and BC = 36 in. Solved Examples Dr. Tarek Hegazy 64

67 Example Example: Dr. Tarek Hegazy 65

68 Buckling of Columns - Slender columns under elastic compression buckle when the load exceeds a critical value. - Buckling causes column instability. - Short stocky columns do not buckle. - W need to study the relation between P,, and shape of buckled column. P P - Analysis (Euler ): υ υ P M M P. υ = 0 Recall, M = d 2 υ EI dx 2 Then, d 2 υ P. υ = 0 dx 2 EI Equation of Elastic Curve: υ = C1 Sin [(P/EI) 0.5. x] C2 Cos [(P/EI) 0.5. x] υ υ= 0 at x = L υ= 0 at x = 0 or when, Sin [(P/EI) 0.5. L] = 0 C2 = 0 or when, (P/EI) 0.5. L = π, 2π,. x 4 Dr. Tarek Hegazy 66

69 Analysis: Maximum axial load before buckling: P/A should be within allowable stresses. Smaller of the two directions x & y. A Put, r = I / A = radius of gyration OR (L/r) 2 Note that L/r is the Slenderness Ratio used to classify columns as long, intermediate, or short. Effect of Column Supports: ; = Solved Examples Dr. Tarek Hegazy 67