285 The motor speed was approximated being loaded to these torque limits, and this was used to estimate  the maximum low and high-speed output RPM and power-out before either clutch would be expected to begin slipping. If a clutch should start to slip, th e output power would drop dramatically because the  slipping friction of the plates would provide a much lower output torque. This could also damage the clutch because it is not designed to slip for long pe riods, or to dissipate the resulting heat. Currently, the  power to the drive motors of the Chariot control sy stem should not provide eno ugh current to the motors  to exceed the clutches’ torque rating.  Gears

The gear teeth pitch and face widths were select ed based on the anticipated operational loads and  speeds. Traditional formulas were used to calculate gear tooth stresses, and to help decide the required gear pitch and face width. The width of the smallest of the two meshing gears at all three gear meshing  planes is 1.5 mm (0.06 in) larger, which allows the gears to tolerate some axial misalignment due to dimensional tolerance stack-ups during assembly. The smaller gear was selected to have the larger width in order to keep the overall weight minimized, and to bolster wear duration of the faster turning gear. For  example, while the large gear #1 has a face width of  only 6.4 mm (0.25 in), which is enough from a gear  tooth-strength standpoint, both motor pinions have a face  width of 7.9 mm (0.31 in) to ensure that, after  final assembly, the entire 6.4-mm (0.25-in) width of gear #1 is engaged. Similarly, gear #2 has a larger  face width than gear #3, and gear #4 is larger than gear #5. Also, both gear #1 and #3 are large enough to be lightened-up by the addition of a pattern of holes, again in an attempt to minimize the weight of the final product. The gears are designed to helicopter-grade quality, or an American Gear Manufacturers Association (AGMA) gear tolerance Class 12.    Casing Design

It was a design goal for the case to shroud the gears as tightly as possible to minimize both the size and weight of the final product. It should be noted again th at for compatibility, the same motors as used in the  original gearbox were also used in the second generation gearbox. Due to time constraints, commercial off-the-shelf clutches were used that were very sim ilar in size and capability to those in the original  gearbox. Although the main drive gears were sized and optimized for minimum weight, the overall gearbox size was largely dict ated by the size of the motors and clut ches. Slightly taller than the original,  the new gearbox is not nearly as broad, and is theref ore smaller and lighter. Considerable effort went into  fine-tuning the gearbox design so that the mounted height of the moto rs would only be slightly above the  mounting plate (Fig. 8).    A stacked-component approach was used in the desi gn of the case. This helped keep the size profile  minimized because it allowed the top gear plane, with the two motor pinions and large gear, to be  shrouded closely by way of the Top Cover before transit ioning to the majority smaller profile required for  the main body which encompasses the rest of the gears and both electric clutches (Fig. 6). A total of five parts make up the complete case:  The Upper Cover, Transition Case, Deep Case, Short Case, and Lower Cover. Having a multi-part casing also redu ced the machining risk on each separate part. The fit  between all parts was kept flat (planar) for machining simplicity.  There was no seal designed between the casing interf aces. Like the original gearbox, no oil containment  was required because grease lubrication was planned. However, a full perimeter lip was designed into the  face-fits between each casing part. This serves to hol d the required relative positioning between each of  the casing parts for shaft and bearing alignments, and to make assembly easier. It also serves as a  partial seal to help keep dirt out, and extraneous grease in. The final wall thicknesses were determined in part through the limited use of finite element analysis. The analysis indicated that the casing thicknesses could have been reduced further, but 3.18-mm (0.125-in)  thickness for most of the walls was determined  to be the minimum from a standpoint of machini ng comfort, stiffness, and warpage concerns.  

NASA/CP-2010-216272