Transactions

502 TRANSACTIONS OF TH E A.S.M.E. AUGUST, 1941
stresses must be conservative and material m ust be high-strength
steel. A factor of safety of approximately 12 to 13, based on the
ultimate strength of the steel, is recommended.
The wings of the saddle type of connecting rod, through which
the wristpin bolts pass, are partly held at their sides. However,
to simplify the method of calculating the depth required, it is
convenient to consider them as simple cantilever beams with the
bolt load as the applied load. The length of the cantilever beam
is the distance y shown in Fig. 1. Bending stress at the junction
of the cantilever and the main shank of the connecting rod should
not be more than the ultimate strength divided by about 10.
For the main-column section of the connecting rod, three
cross-sectional types are used: I-beam, round, and rectangular.
The minimum area of any cross section should be such that
the compression stress per square inch of area from peak cylinder
load is not more than the ultimate strength of the steel divided by
approximately 7.
In addition to direct stress on the minimum cross-sectional area,
the connecting rod is a column and should be designed for safe
stress as a column. When bending in the plane of the crankshaft
center line, both ends of the column are considered fixed,
with the column length as the distance from the crankpin center
line to the wristpin center line, Fig. 3(6). Rankine’s formula for
short columns has been found to be the most satisfactory, as
very rarely is the l/r ratio of connecting rods greater than 120.
When bending in a plane 90 deg to the crankshaft center line,
the ends of the rod are considered free, Fig. 3(a). Thus, in the
latter direction, the cross section of the rod must be designed with
a greater radius of gyration r than in the other direction. Rankine’s
formula for short columns is
Sc = (PM )[1 + K l/rY).............................[6]
where P = total thrust load on rod, lb; A = cross-sectional area
at mid-point of rod, sq in.; k = 0.00004 for columns of fixed
ends; k = 0.00016 for columns with free ends; I = length of rod
from center line of crankpin to center line of wristpin, in.; r =
radius of gyration of mid-point cross section in the direction that
bending is being considered, in.; and Sc = column stress, psi.
The latter value should not exceed the ultimate tensile strength of
the material divided by a factor of safety 6.5.
Using this method of design, it is quite obvious that the maximum
stiffness and strength, with the minimum weight of material,
is obtained by the use of I-beam cross-sectional shapes with the
long axis of the I-beam in the plane 90 deg to the crankshaft
center line. On some types and sizes of engines, where a low rate
of production does not warrant the expense of dies for forging
the I-beam section, it is more economical to use either the round
or rectangular shape.
The crankpin end of the connecting rod must be made stiff to
prevent bending of the bolts holding the cap, and to prevent deflection
of the crankpin-bearing shell. Most failures of crankpinbearing
bolts and some failures of crankpin bearings may be
attributed to the lack of stiffness in this end of the connecting
rod. It is impractical to state the maximum allowable deflection
in this end of the rod as this must be judged mostly by experience
obtained from previous designs. The ideal deflection would, of
•course, be zero which is obviously impossible.
The flanges of the I-beam of the connecting-rod shank should
be extended right to the bearing-shell backing to help stiffen the
support for the bearing, as shown in Fig. 1. The angular cross
section on approximately a radius from the center line of a
crankpin bearing to the curve where the rod flares out to form
seats for the bolts should also be of I-beam cross section for greatest
rigidity with minimum weight. The bolts should be fitted
a t the joint between the rod foot and the connecting-rod-bearing
cap, or heavy closely fitted dowels should be used.
F i g . 4
A W e a k C o n n e c t i n g - R o d F o o t
A lightweight design which caused considerable difficulty in
broken bolts, due to deflection of the rod foot, is shown in Fig.
4. A comparison of Fig. 4 with Fig. 1 will show the great gain
in stiffness in the latter design with only a slight increase in
weight, the increase in stiffness being in the ratio of approximately
4.5 to 1. The type of design shown in Fig. 1, which is now
more universally used, has eliminated crank-bolt breakage previously
caused by bending, due to deflection of the crank end of the
connecting rod. It can also be shown mathematically that the
centrifugal force of the rotating end of the connecting rod applied
outward, when crank and connecting-rod center lines are approximately
90 deg with each other, that the deflection of the design,
shown in Fig. 4, will cause a bending stress in the bolts of approximately
34,000 psi. Adding to this the direct tensile stresses on
the bolts, we find that the actual working stress is practically at
the endurance limit of the material. This accounts for the
failures.
C r a n k - B e a r in g B o l t s
In four-cycle engines, the crank-bearing bolts must withstand