176 size the members and select materials, finishes, and clearances to  better reflect tribological best practices .  Early prototypes  experienced  a variety of failures at this interface, both with roller bores wearing and  enlarging to unacceptable sizes and with crank pins wearing severely and, in some cases,  break ing off  completely as a result of the drastic reduction to their cross section by wear .  

A redesign attempted to reduce the bearing load associated with the highest percussion output level to  below 3.45 MPa ( 500 psi ). Very often such plain bearing pairs use a hard shaft in a bearing that is made  from a copper -based,  chemically dissimilar metal and usuall y the  bearing is softer and will preferentially  wear . Such systems are good when a number of common assumptions are true that do not apply to this  mechanism – specifically that the parts are serviceable and that petroleum- based lubricants will be used.  The strategy adopted for this design was to si mply create the longest lasting (slow est wearing) possible  design by making both parts as  hard as possible, reasonably similar in hardness, and also by reducing the  starting clearance between the parts to the smal lest comfortable value, to minimize impact  loads  between  the roller and shaft when the driver plate changes direction. The final design increased the overall size of  the parts from 3.175 mm (.125 in)  shaft to 4.7625 mm (.1875 in)  and 6.35 mm (.25 in)  roller ball size to  9.525 mm (.375 in) . The shaft was made significantly harder by the addition of a 75 µm (.003 in) thick nitride  case approximately 65 to 68 HRC so as to balance the high hardness of the ball, which would be made  from a Grade 10 440C ball, 58 HRC minimum . Shaft to bore clearances of approximately 7.6 µm ( .0003 in)  on diameter were achieved by manufacturing balls with several graded bore sizes . The shaft and bore both  had 0.2 µm (8 µin) surface finishes specified . As before, the  bearing was filled with Castrol Micronic  Braycote 600EF grease and disassembly at the end of a 3x  life test showed some grease was still present  between the parts.  

Challenge: Hammer bushing wear and retention   Development prototypes failed to achieve acceptable levels of wear  life in the hammer bushings and a  threaded retainer that was used to capture the upper bushing catastrophically failed shortly after 1x life.  Many brute force attempts were made to improve the retainer –  to pin it in place, to drill point it and drive  set screws through the hammer shaft into the retainer, and to peen it into place . LS-Dyna analysis showed  that the loads in this region were very high and, more importantly, the back of the hammer shaft experiences  a large load reversal w hen the traveling compression wave from the impact encounters the end of the part  and starts a tensile wave in the reverse direction. The small volume of the region did not lend itself to robust  fastening features, perhaps besides welding, which was not at tempted .  

The solution was to eliminate any sort of bushing retainer altogether and capture the bushing in a blind  groove. To enable this simplification, a change in bushing material had to happen at the same time, so that  the bushing could be designed to be installed into this groove. A split bushing made from DuPont Vespel®  SP-3 (MoS 2 impregnated) was sufficiently flexible that it could be wound in on itself and inserted past the  shoulder of the groove (see Figure 8) and as it turned out the Vespel® SP-3 was also slightly slower wearing  than the prior metal/polymer composite bushing, a Glacier Garlock Bearings (now “GGB”) DP4®. A key  take- away for other percussion mechanisms is to reduce or eliminate small parts whenever possible that  might not easily be able to grow to accommodate high loads , and particularly to expect high reversing l oads  at the back end of the hammer or striker.  

Challenge: High drag during cold operation   The initial flight design called for 10% to 15% grease fill in all ball bearings . All the bearings have a phenolic  retainer impregnated with Castrol Brayco 815Z oil and use Castrol Braycote 600EF grease. Although much  early life testing was done on prototype units, including some cold testing, none had gone below -40°C .  The first tests at the min imum  qualification temperature of -70°C showed that the mechanism cou ld not  reliably start from cold and the actuator would sometimes stall at the 10 A current limit, equivalent to   0.75 N·m [6.6 lbf·in] of torque at the actuator output shaft .