Rose does the books and knows exactly how much I spend, most of the time. Lots and lots of money has been spent since the Pendine race with no results and she knows this, too. She hasn't said anything negative and I am staying within my $500 a month race budget, sort of. She rode pillion on our first date on the back of a BSA A-65 Spitfire. That was 40 years ago. Maybe she figures bikes and husband are a package deal.
The first stage in the build design is done and it used the filling-emptying model with ten point cam data. There is a procedure for optimizing bore and stroke. That cannot be changed due to budget considerations so I did not look at it. The first thing looked at was inlet port minimum area. The formula is on Page 250 in the users manual and it is (RPM x Stroke x Bore Squared) / 190,000. The minimum area is 2 x (8,500 x 2.677 x 3.8**2) / 190,000 = 1.73 square inches.
The minimum port area is between the valve seat and the hole for the valve stem. It is 1.89 square inches for the two ports. The engine has adequate intake port minimum area. There are all sorts of problems if this area is too big so it will be left as is. The intake valve sizes will stay at 2mm larger than standard. There is flow data for this size valve and an old and smaller manifold so I had to estimate flow with the new bigger manifold. Kibblewhite is asked to do a flow test with the bigger manifold so I can rerun the simulations using measured values.
The second thing examined was the exhaust valve size. There is a measured flow vs lift curve for 26mm valves. The low lift flows for bigger valves are extrapolated based on increases in curtain periphery. The high lift flows are based on the proportional increase in valve head area. The flows at intermediate lifts are interpolated between these. This represents bigger valves, seat diameter enlargement, and some grinding on the ports.
All cam timing simulations compared the three different valve sizes, 26, 27, and 28 mm. The compression losses due to combustion chamber modifications were also considered. This is 11.7 to 1 with standard exhaust valves, 11.5 to 1 with 1mm larger valves, and 11 to 1 with the 28mm valves. The 27 mm exhaust valves consistently gave better performance in the simulations. The flow data estimates are sent to Kibblewhite. They will be asked if these flows appear to be obtainable with using the existing valve seats, enlarged a tad, with 27mm valves. This will be done if it is possible. The compression loss associated with bigger seats will be too much and the 26 mm valves will be the best option if this occurs.
Cam timing was the third thing considered. Sensitivity analyses were made to see how performance changed with variations. Lift was not a significant factor as long as it was over .370 to .380 inches. Duration was not critical as long as it did not get too far from the optimized values of 273 for intake and 284 for exhaust at seat-to-seat (.006 lift). Cam timing was most critical with 102 for intake centerline and 115 for exhaust.
The all-purpose cams I have used for years have .380 lift. This is adequate. Seat to seat duration extrapolated from .050 timing values is 291.6 for both. These cams, with intake and exhaust lobe centers at 101.5 and 114 give good simulated horsepower. These cams will be used. Kibblewhite is asked to send me digital lobe profiles so I can do more complex wave simulation modeling. They are also asked to check spring tension and valve train harmonics with these cams.
Last, rpm at peak power was looked at. Cam timing and valve size could be varied to move the horsepower peak up from 8,500 rpm to 9,500 or higher. Peak horsepower did not increase. Torque decreased and this was offset by the higher rpm to produce the same power. Peak power target rpm is kept at 8,500. There is less chance of wearing out and breaking stuff at that rpm.
That is it for now. The head, cams, etc were sent out yesterday.