AGMA 09FTM19-2009 pdf download.AGMA Technical Paper The Effect of Gearbox Architecture on Wind Turbine Enclosure Size By C.D. Schultz, Beyta Gear Service.
Design constraints
An experienced gearbox designer has usually developed a set of guiding principles to speed his or her work. The author has spent much of his career designing special, one-off gearboxes where a conservative design philosophy is required out of respect for a lack of qualification testing and development time. The constraints adopted for this paper are reflective of that experience and the author recognizes that other designers may disagree with the limits he has established. The reasons for each of these constraints is discussed in the following paragraphs.
Minimum number of pinion teeth
The choice of 18 for a minimum number of pinion teeth was made based upon maximizing the tooth strength, achieving a minimum profile contact ratio of 1.30, and reducing the grind cycle time. [Form grinding cycle times are a function of the number of teeth, stock ailowance. and face width.) Having designed parallel axis gear sets with as few as 3 pinion teeth and as many as 42 pinion teeth, 18 is a good minimum to avoid hobbing issues [undercutting, problems with start of active profile overlapping the top of the fillet] while still providing an acceptable profile contact ratio.
Maximum face width/pinion pitch diameter
ratio
As gear capacity and cost tend to follow a volume function, pay careful attention to the ‘FD squared’ principle (where F is the face width and D is the pinion pitch diameter). It was not unheard of. back in the 1960s and 1 970s, to have a face diameter ratio of up to 2.00 in through hardened industrial gearboxes. As the service hours accumulated on these long thin pinions it became apparent that torsional deflection adversely effected the life of these drives. In Later design work we have had the opportunity to see the beneficial effects of reducing the FID ratio to the 1.00/1.25 range and have avoided using a higher value ever since.
Minimum face contact ratio. M.
If helical geometry is to be fully effective, a minimum face contact ratio of 1.00 per helix is needed. The adjustments in the gear rating formulas to account for M values of less than 1.00 have limited testing behind them so they should be avoided. Once the complications of thrust and overturning moment are introduced to the bearing evaluation process, it seems prudent to insure that the gears will enjoy the full benefits of helical load sharing.
Number of planets
Figure 1 shows the geometry behind my limits on the number of planets. We recognize that nonstandard geometries can allow some adjustment to these ratio limits but find them to be good guidelines for general design. As ratings are all about “power per mesh” we have chosen to use the maximum number of planets wherever possible.
Maximum individual mesh ratio
The FD squared” principle referenced earlier plays a big part in the decision to limit individual mesh ratios to less than 6.5:1 except in the case of a single stage 10:1 double helical gear set. That exception serves as an excellent illustration of how rotating mass increases very rapidly as set ratio goes up. see Figure 2, case A.
Radial timing
As mentioned above, rating calculations are based upon power per mesh. When multiple meshes are used to share the load it becomes incumbent upon the designer to insure that load sharing is uniform or that the drive train can accommodate the anticipated degree of inequality. Our experience with industrial divided power path drives makes us very skeptical that uniform distribution ever occurs and the highly variable nature of the loads in wind turbines further increases my discomfort. For this reason we have limited the designs in this paper to those which do not require radial timing or load sharing adjustment outside the planetary stages.
Planet load sharina
Load sharing within planetary stages is widely eod within the gear design community.AGMA 09FTM19 pdf download.