Bending

In applied mechanics, bending (also known as flexure) characterizes the behavior of a slender structural element subjected to an external load applied perpendicularly to a longitudinal axis of the element.
The structural element is assumed to be such that at least one of its dimensions is a small fraction, typically 1/10 or less, of the other two. When the length is considerably longer than the width and the thickness, the element is called a beam. For example, a closet rod sagging under the weight of clothes on clothes hangers is an example of a beam experiencing bending. On the other hand, a shell is a structure of any geometric form where the length and the width are of the same order of magnitude but the thickness of the structure (known as the 'wall') is considerably smaller. A large diameter, but thin-walled, short tube supported at its ends and loaded laterally is an example of a shell experiencing bending.
In the absence of a qualifier, the term "bending" is ambiguous because bending can occur locally in all objects. Therefore, to make the usage of the term more precise, engineers refer to a specific object such as; the "bending of rods", the "bending of beams", the "bending of plates", the "bending of shells" and so on.
Quasi-static bending of beams.
A beam deforms and stresses develop inside it when a transverse load is applied on it. In the quasi-static case, the amount of bending deflection and the stresses that develop are assumed not to change over time. In a horizontal beam supported at the ends and loaded downwards in the middle, the material at the over-side of the beam is compressed while the material at the underside is stretched. There are two forms of internal stresses caused by lateral loads:
These last two forces form a couple or moment as they are equal in magnitude and opposite in direction. This bending moment resists the sagging deformation characteristic of a beam experiencing bending. The stress distribution in a beam can be predicted quite accurately when some simplifying assumptions are used.
Euler–Bernoulli bending theory.
In the Euler–Bernoulli theory of slender beams, a major assumption is that 'plane sections remain plane'. In other words, any deformation due to shear across the section is not accounted for (no shear deformation). Also, this linear distribution is only applicable if the maximum stress is less than the yield stress of the material. For stresses that exceed yield, refer to article plastic bending. At yield, the maximum stress experienced in the section (at the furthest points from the neutral axis of the beam) is defined as the flexural strength.
Consider beams where the following are true: 
In this case, the equation describing beam deflection (formula_1) can be approximated as:
where the second derivative of its deflected shape with respect to formula_3 is interpreted as its curvature, formula_4 is the Young's modulus, formula_5 is the area moment of inertia of the cross-section, and formula_6 is the internal bending moment in the beam.
If, in addition, the beam is homogeneous along its length as well, and not tapered (i.e. constant cross section), and deflects under an applied transverse load formula_7, it can be shown that:
This is the Euler–Bernoulli equation for beam bending.
After a solution for the displacement of the beam has been obtained, the bending moment (formula_6) and shear force (formula_10) in the beam can be calculated using the relations
Simple beam bending is often analyzed with the Euler–Bernoulli beam equation. The conditions for using simple bending theory are:
Compressive and tensile forces develop in the direction of the beam axis under bending loads. These forces induce stresses on the beam. The maximum compressive stress is found at the uppermost edge of the beam while the maximum tensile stress is located at the lower edge of the beam. Since the stresses between these two opposing maxima vary linearly, there therefore exists a point on the linear path between them where there is no bending stress. The locus of these points is the neutral axis. Because of this area with no stress and the adjacent areas with low stress, using uniform cross section beams in bending is not a particularly efficient means of supporting a load as it does not use the full capacity of the beam until it is on the brink of collapse. Wide-flange beams (-beams) and truss girders effectively address this inefficiency as they minimize the amount of material in this under-stressed region.
The classic formula for determining the bending stress in a beam under simple bending is:
where
Extensions of Euler-Bernoulli beam bending theory.
Plastic bending.
The equation formula_19 is valid only when the stress at the extreme fiber (i.e., the portion of the beam farthest from the neutral axis) is below the yield stress of the material from which it is constructed. At higher loadings the stress distribution becomes non-linear, and ductile materials will eventually enter a "plastic hinge" state where the magnitude of the stress is equal to the yield stress everywhere in the beam, with a discontinuity at the neutral axis where the stress changes from tensile to compressive. This plastic hinge state is typically used as a limit state in the design of steel structures.
Complex or asymmetrical bending.
The equation above is only valid if the cross-section is symmetrical. For homogeneous beams with asymmetrical sections, the maximum bending stress in the beam is given by
where formula_21 are the coordinates of a point on the cross section at which the stress is to be determined as shown to the right, formula_22 and formula_14 are the bending moments about the y and z centroid axes, formula_24 and formula_16 are the second moments of area (distinct from moments of inertia) about the y and z axes, and formula_26 is the product of moments of area. Using this equation it is possible to calculate the bending stress at any point on the beam cross section regardless of moment orientation or cross-sectional shape. Note that formula_27 do not change from one point to another on the cross section.
Large bending deformation.
For large deformations of the body, the stress in the cross-section is calculated using an extended version of this formula. First the following assumptions must be made:
Large bending considerations should be implemented when the bending radius formula_28 is smaller than ten section heights h:
With those assumptions the stress in large bending is calculated as:
where
The general solution of the above equation is
where formula_37 are constants and
formula_38
Timoshenko–Rayleigh theory.
In 1877, Rayleigh proposed an improvement to the dynamic Euler–Bernoulli beam theory by including the effect of rotational inertia of the cross-section of the beam. Timoshenko improved upon that theory in 1922 by adding the effect of shear into the beam equation. Shear deformations of the normal to the mid-surface of the beam are allowed in the Timoshenko–Rayleigh theory.
The equation for the bending of a linear elastic, isotropic, homogeneous beam of constant cross-section under these assumptions is
where formula_40 is the polar moment of inertia of the cross-section, formula_41 is the mass per unit length of the beam, formula_28 is the density of the beam, formula_32 is the cross-sectional area, formula_44 is the shear modulus, and formula_45 is a shear correction factor. For materials with Poisson's ratios (formula_46) close to 0.3, the shear correction factor are approximately
Free vibrations.
For free, harmonic vibrations the Timoshenko–Rayleigh equations take the form
This equation can be solved by noting that all the derivatives of formula_1 must have the same form to cancel out and hence as solution of the form formula_50 may be expected. This observation leads to the characteristic equation
The solutions of this quartic equation are
where
The general solution of the Timoshenko-Rayleigh beam equation for free vibrations can then be written as
Quasistatic bending of plates.
The defining feature of beams is that one of the dimensions is much "larger" than the other two. A structure is called a plate when it is flat and one of its dimensions is much "smaller" than the other two. There are several theories that attempt to describe the deformation and stress in a plate under applied loads two of which have been used widely. These are
Kirchhoff–Love theory of plates.
The assumptions of Kirchhoff–Love theory are
These assumptions imply that
where formula_56 is the displacement of a point in the plate and formula_57 is the displacement of the mid-surface.
The strain-displacement relations are
The equilibrium equations are
where formula_7 is an applied load normal to the surface of the plate.
In terms of displacements, the equilibrium equations for an isotropic, linear elastic plate in the absence of external load can be written as
In direct tensor notation,
Mindlin–Reissner theory of plates.
The special assumption of this theory is that normals to the mid-surface remain straight and inextensible but not necessarily normal to the mid-surface after deformation. The displacements of the plate are given by
where formula_64 are the rotations of the normal.
The strain-displacement relations that result from these assumptions are
where formula_66 is a shear correction factor.
The equilibrium equations are
where
Dynamic bending of plates.
Dynamics of thin Kirchhoff plates.
The dynamic theory of plates determines the propagation of waves in the plates, and the study of standing waves and vibration modes. The equations that govern the dynamic bending of Kirchhoff plates are
where, for a plate with density formula_70,
and 
The figures below show some vibrational modes of a circular plate.