Let's examine some of the contents and see what can be learned from it.
Kevin has provided links to some pages on crankshafts here:
I quickly retyped some of it to make it easier to read, (the copies were difficult to read) and to highlight some points I want to observe.
Most designers plot diagrams of crankpin and main journal loads based on indicator diagrams and calculated inertia forces using rigid-body assumptions. The results are translated into nominal working stresses, which are held to limits dictated by experience with similar designs. As previously indicated such procedures are basically an application of the theory of similitude since accurate computation of loads and stresses is not possible in such a complex system.
Today we can accurately simulate the flexible bodies with dynamic loads (not useless static rigid bodies). This is a critical difference as this is what is required to deal with harmonic vibrations within a crankshaft as this is what causes them to break. The author could have learned more in a few days with a modern CAD and FEA software than what is printed in this book.
250-275 Brinell (that’s about 24-28 Rockwell C)
Bad advice, a racing crank is more like a minimum of 38 C, I have made them much higher.
Qualitatively speaking, the loads on a crankshaft result in stresses due to bending, torsion, and shear throughout its entire length. The complex geometry involved would make accurate stress computations impossible even if the loads were accurately known. In spite of these difficulties, however, much has been done toward rationalization of crankshaft design, largely by means of experimental stress analysis.
Again, we have very accurate and convenient tools to do this now. It seems that our present times tech was on the authors wish list.
Results of Stress Analysis
The most useful information in regards to crankshaft structural design comes from experimental studies of stress distribution in typical crankshaft samples or models subjected to arbitrary loads under laboratory conditions. Figures 11-35a shows the enormous effect of fillet radius on the maximum stress in a straight shaft and a simple crank. Given reasonably well-chosen ratios of major dimensions, fillet radius is the most important factor in crankshaft strength. The larger the fillet radius the better, provided adequate space is available for the required bearing lengths. Where bearing space is at a premium, undercut fillet radius are better than inadequate conventional ones. Fillets with non-circular contours are slightly better than fillets with a fixed radius but are not usually practicable except for large crankshafts.
Obvious to a 1st year vocational school machinist getting better grades than C.
An image shows the imagined stress concentration in a crankshaft, they are not even remotely close to reality.
the stress was measured with an extensometer
This is a device that is intended to be used to measure stretching of objects, it was really only useful/accurate to use on flat surfaces, it was not a reliable way to measure something shaped like a crankshaft. Besides it tells nothing about the internal stresses. That's why the hand drawn stress images are so far from reality, they were just someones imagination.
the stress can be several times higher than that calculated by the simple beam and shaft assumptions. Fillet stresses up to five times the nominal torsional stress and three times the nominal bending stress.
To make informed decisions about the design of a stressed object, you need to accurately simulate the loads and see where the stresses are and how high they are and how they are distributed. The author is explaining that the best estimate they have is within multiples of reality. Today we can come within a few %, closer if you have the need and time.
The bevel shown in figure b was probably detrimental to strength at least in the all of the bored shafts,,,
This is the typical taper in a rod throw seen from the side of a crank, it might be concave, convex and have contoured shape today. Today we tune the stiffness of the crank with varying shapes in that area depending on many variables of the application.
Where stresses are not considered to be high, crank cheeks are allowed to remain in the as-forged or as-cast condition.
OK, good luck selling that. If China turned out a part with unfinished cheeks they would be ridiculed.
In order to reduce wear crank pins and journals are hardened by local heat treating of the surfaces either by flame or by electric induction.
Crankshaft design ratios:
Where it is required that the connecting rod be removable through the cylinder bore crank pin to bore ratio is limited to a value of 0.6
Main journal to bore ratios 0.6 to 1.0
NASCAR ~4.2 x 2.0 = 4.76 (well outside the recommended range)
It is remarkable that the range of ratios is about the same for all engines listed.
It is also remarkable that these tend to be rather independent of specific output.
That might have been true back then when engines were at 1/2 HP pr cubic inch and operated at low speeds. We have a lot wider variety now and different needs.
Crank pin diameter should be at least 0.6 times the bore
Main journal diameter should be large than crankpin diameters.
So a NASCAR engine should have a 2.508 rod pin diameter?
And the main 3.0"
They run 45mm rod which is 1.77, way off.
So I think it is fair to say that most of the stuff on those pages was either obvious or wrong.
Someone wanting to gain knowledge would be much better off to spend the same time looking at modern parts and learning CAD and FEA at a community college.
Main journal length can be as short as 0.30 times the journal diameter when centrifugal loads are counterbalanced.
Wow. we have come a long way.