Page 218 - Failure Analysis Case Studies II

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The lead wires and bridge energising wires were all led through the middle of the shaft to the end
and permanently attached to a custom built junction connector glued to the end of the shaft. In this
way the straingauged shaft could be conveniently handled and other attachments or machine
elements such as nuts, disc brake, lock nut assembly, etc, easily fixed to the shaft in the usual
manner.
After assembly the strain gauge signals under load conditions could be monitored by simply
soldering fine wires of up to 2 m in length to the junction connector and in turn connecting these to
the strain gauge amplifier and digital storage scope facility. Thus it was not necessary to have a slip
ring system or small transmitter telemetry device which would have altered the balance and stress
conditions at 1400 rpm. The shaft was seldom running for more than 15-20 s at a time, and in any
case the rotation dirKtion could be readily reversed to avoid excessive twisting of the wire leads.
Under test conditions the twisting of the wires, even at 1400 rpm, was quite acceptable and the
concept worked elegantly. When the wires became too twisted they were discarded and readily
replaced.
The recessed diameter section was protected with an epoxy glue which was machined smooth
after drying so that the Belvel washers readily slipped over onto the full shaft diameter without
causing any damage, or spurious strain reading, to the strain gauges themselves. The integrity of
the gauges and wiring was checked at all stages of development. Since the strain gauge lead wires
were permanently fixed to the junction connector at the end of the shaft, any stress condition of the
shaft (e.g. bending or tension) could be selected simply by wiring up the appropriate long twisting
wire (i.e. TW) leads to the strain gauge amplifier.
It should be borne in mind that the raw strains recorded would be appropriate for the reduced
section actually measured (diameter 15.5 mm, and including the central bored hole) and the resultant
stresses determined required modification by a factor (0.825, Le. reduced) if the stresses were to
refer to the original shaft diameter at the thread roots (diameter 16.5 mm). It should also be borne
in mind that no allowance has been made for the stress concentration effect of the threads which
would increase the localised stress by a factor of typically between 2 [5] and 4 [6].

2.2. Lathe calibration
To assess the performance of the strain gauged wormshaft system under rotation conditions, as
well as to calibrate the facility, it was mounted on a lathe which could be set to run at 1500 rpm
(close to the test bench rotational speed).
Firstly the TW gauge leads were connected to check sequentially for tension, torsion and bending
stresses, but at zero angular speed and for each of these configurations performance was satisfactory.
It was then necessary to assess performance under rotational conditions. The wormshaft was
mounted in a lathe and the main bearing fixed to the shaft at a distance of 75 mm from the strain
gauges. Load was applied controllably, through a 5 kN load cell mounted on the tool post, to the
bearing, to simulate known bending conditions under rotation. A schematic diagram of the lathe
calibration system is shown in Fig. 3(b).
With this system the strain gauges performance in bending can be checked and in effect “cali-
brated”. The effective conversion at the voltage selections used on the digital voltmeter was 2.2
microstrain per millivolt and the behaviour was linear and this was also used for the bench tests at
Koeberg.

2.3. Test bench programme
The test bench programme at Koeberg needed to measure the stress (inferred from the strains)
that might arise from the disc brake loading and to distinguish between bending, tension and torsion
components. To facilitate this the trip switch torque limiter device was overridden (i.e. disengaged)
so the disc brake torque limiter would indeed be loaded against the brake pads. Typical torque
conditions for these tests on the test bench were at a torque of 190k 15 Nm. For completeness, tests
were also conducted (i) with the trip switch engaged (so that the brake did not touch the brake
pads); and (ii) with the whole disc brake torque limiter system removed and the actuator simply
stopped under so called “stall” test conditions. It was expected that under this latter test condition
there would be only low tensile stresses and no bending stresses at the strain gauge, since the disc