Before you can analyse Contact-Force and Contact-Stress, you need to reconfigure the model so that it transfers the forces correctly between the cam and the follower.
Cam-followers are bought 'off-the-shelf' [from catalogues and now the web]. Their selection is based on design loads, with working environment, contamination, lubrication, and reliability also being considered. However, there is a standard procedure the engineer can use to select a cam-follower. If the cam-follower fails, then the difference between the actual and design environment can be reviewed. The working environment can be modified or a different cam-follower can be selected. It is rare that a cam-follower is faulty. Its steel quality, manufacture and heat-treatment are maintained to exacting standards. Changing the manufacturer of an industrial cam-follower will not usually change the performance outcome of the machine.
However, the steel, manufacture, and heat-treatment processes and specifications for cams are confidential to each machine supplier. Each supplier will not usually divulge their standards to competitors.
Automobile cam-shafts are manufactured 100,000, and more, for 'mature' engines. The cams are fully lubricated and wear is minimal for 10,000,000+ rotations.
However, industrial packaging, assembly, or textile machines are manufactured 1 to say 100 times. Sometimes many times more, but often less than 10 times.
Even though a cam is an easier machine element to analyse, it is left to the machine designer to choose the correct steel, manufacturing and heat-treatment processes and method of lubrication, to give the cam an operational-life as required by their customers. Heat-treatment is often sub-contracted, and its quality control is not specified.
As a cam rolls under a loaded cam-follower roller (or the cam-follower rolls over a cam) the cam and the cam-follower become stressed and the cam and cam-follower become slightly flattened, The stresses at and below the contact-surface are quite complex. They are nothing like the stresses that you can calculate from Tension, Bending or Torsion. Hertz was the first to calculate the stress 'field' and it is now frequently termed Hertzian Contact Stress [Reference Hertz (1881) has done the hard work to derive the equations we need to calculate the stress when there is contact between two bodies that are conformal or non-conformal.
[H. R. Hertz, "Über die Berührung fester elastischer Körper (On Contact Between Elastic Bodies)" Journal für die reine und angewandte Mathematik 92, 1881 pp. 156-171].], Hertzian Stress, or Contact-Stress.
MechDesigner calculates the Hertzian Contact Stress. and the Maximum Shear Stress which, as we will see, is quite important. It is convenient to consider two types of Hertzian-Contact:
Line-Contact : between two parallel cylinders, or between a cylinder and plate. This is equivalent to a cylindrical cam-follower and a 2D-Cam, for example.
Elliptical-Contact : between bodies with more complex shapes. Their shapes are usually defined by their radii-of-curvature in orthogonal directions. When the principal radii and their direction are different, then the calculations are much more difficult. However, we can give very good approximations when the bodies are similar to a barrel cam-follower rolling over a cam-surface.
Line Contact : there are only two radii at the contact that we must know:
When the Cam and Cam-Follower lightly 'touch' each other, the contact is a 'line'. As we increase the contact force, the cam and follower deflect and the contact becomes a narrow rectangle across the cam and cam-follower.
Parameters to calculate Hertzian Contact-Stress for Line-Contact.
Equivalent Elastic Modulus:
Half-width in direction of rolling:
Mean Hertzian Stress:
Maximum Hertzian Stress:
Substituting for 'b' gives:
Pmax is at the middle of the contact patch, where the contact of a very light-load is initially made.
Contact Stresses for Line Contact
Poisson's Ratio Shear Stress
The image to the left shows principal stresses [all compressive] relative to the Maximum Hertzian Contact Stress, Pmax , directly below the centre-line of the contact.
Horizontal-Axis : Scale: 1.0 = Pmax : Vertical-axis : Scale: b = half contact half-width
σz : Stress in the vertical direction
σx : Stress in the horizontal direction, in direction of rolling
σy : Stress in the horizontal direction, in direction that is across the width of contact
Ԏ45 : Maximum Shear Stress that results from the maximum difference between the principal stresses.
Maximum Value of Shear-Stress, Ԏmax [Line Contact]
Ԏ45 = 0.3×Pmax
Depth of Maximum Shear-Stress [Line-Contact]
= 0.78 × b
The image to the left shows that the stresses, σz, σx and σy are compressive. Because the contact-stress drops to very low values near to the contact, the surrounding material (the hinterland) does not move away to accommodate the expansion because the hinterland is hardly stressed. Because of symmetry of loading, it can be shown that the x, y and z, stresses are 'Principal Stresses'.
Mohr's Circles for the Hertzian Contact Stress at a depth of 0.78b, which is the depth of the Maximum Shear Stress.
This is another way of representing the stress at the depth of Maximum Shear Stress.
Principal Shear Stresses, σx , σy , σz
The image to the left shows the Mohr's Circles for the Principal Stresses, , σx , σy , and σz , relative to the Maximum Contact Stress, Pmax at the urface at a depth 0.78b below the surface.
σx @o.78b ÷ Pmax = –0.32
σy @o.78b ÷ Pmax = –0.15
σz @0.78b ÷ Pmax = –0.75
Maximum Shear Stress, Ԏmax
From the scale of the Mohr's Circle, we can see that the Maximum Shear Stress[vertical axis] is one half of the maximum difference between Principal Stresses:
[σz – σy ] / 2 = [ –0.75 – (–0.15) ] / 2 = –0.6 / 2 = –0.3
Ԏmax = 0.3×Pmax
In many cases, surface roughness, friction, lubrication, thermal effects, and residual stresses will result in conditions that invalidate the exact results from Hertzian analysis. Consequently, the stresses computed according to Hertz’s analysis are often regarded as guidelines that are correlated with experimental failure tests to find allowable stress limits.
Orthogonal Shear Stress moving through the contact zone - image to left.
The Hertzian Contact Stress contact-stresses lead to Octahedral shear-stress and Orthogonal shear-stresses under the surface at the contact. The image to the left schematically illustrates how a cylinder rolling over a flat body in the absence of friction develops subsurface stresses. The stress created just below the centre of the contact is the maximum shear stress (octahedral). We have calculated this in some detail using the analysis above. It occurs at 45° to the contact surface.
Orthogonal shear stresses which are oriented parallel and perpendicular to the contact surface are created in-front of and behind the point of contact. The leading orthogonal stress has the opposite sign of the trailing stress. The magnitudes of the orthogonal stresses are always lower than the magnitude of the octahedral shear stress. However, the range of the orthogonal stresses is higher than the octahedral stress and is thought to be a more potent contributor to the development of contact fatigue damage. The Orthogonal Shear Stress is believed to be significant with respect to 'sub-surface Fatigue Cracks'.
The Orthogonal Shear Stress is greatest at approximately:
Shear Stress and Traction with Rolling Friction - image above
Many elements that are in contact will slide and roll as they pass over each other. For example, gear-teeth, and flat-faced cam-followers sliding against a cam.
There is always a coefficient of friction and thus a friction force. This results in tangential, normal and shear stresses that are superimposed on the stress caused by the normal Hertzian Contact Stress. These stresses are illustrated schematically in the image to the above.