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If the shoe fits......
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I’ll begin this by first saying that there aren’t many forms of internal combustion engines, or any form of motorsports where the engine development programs have been any more challenging than in NHRA Pro Stock competition. I fully realize that this statement will draw fire, but regardless of the fact that these engines only power a vehicle a quarter mile at a time, and this is an "unlimited heads up" class, the rules governing the configuration of these engines make the design and construction of a competitive power plant very difficult. All the rules governing other "unlimited" type engines are considerably less restrictive in all other forms of competition, so I think the best place to start is with the rules.
A quick run down on NHRA"s Pro Stock engine rules will "read" like this: Engine.90 degree automotive-type V8, reciprocating, normally aspirated, single distributor, internal combustion engine. Block can be any material or manufacture, single cam. Max. bore spacing of 4.90", and max. total displacement of 500cid. Two automotive type 4 bbl. carbs, with any internal mods. No split carbs, no fuel injection (yet). Intake manifolds are unlimited in configuration, as long as they’ll fit under a single opening hood scoop with a max. height of 11". Heads must have a max. of two valves, one spark plug, and they must be "castings" with a manufacturer’s logo and pt. #. Now that’s about it, other than the fact that the fuel must be gasoline, and all components such as balancer, and flywheel must be approved. There are also fuel system safety rules, ground clearance, etc.
So, here’s the deal. We need an engine that will propel a 2350lb., front engine car down a quarter mile in less than 7.00 sec. at 200 mph. How to do?? First we study the aerodynamics of the car, the max. allowable hood scoop height (dictated by driver vision and aero studies. The chassis configuration is next with finite stress analysis a must to determine flex and resultant suspension reaction abilities. The transmission type is another variable, although, we all are using clutchless 5 speed square cut boxes currently. And then there’s the clutch. I don’t care how much power you can produce, if you can’t make the clutch slide the correct amount on launch, and each gear change, you won’t qualify. Today the 16 qualifiers are typically running within a .05 sec. separation, and there are 20 more that were within .10 sec. that will either stay and watch or head for home. Pretty competitive, relative to any other form of motorsports competition, especially when you consider that they don’t have additional laps to make up for driver error.
As I’ve never been high on running excessive rpm, we’re going to attempt to build an engine where I dictate the torque curve. Now this is a different approach for most people, since they usually adapt the car to whatever the engine "dictates", but engines are stupid, and my way is considerably less time consuming and less costly. Typically Pro cars are running almost 10,000 rpm which is really incredible when you consider the displacement and the fact that they are single cam push rod engines. The fact that they’re carbureted has absolutely no effect on their power when compared to any known digital fuel injection, assuming you’ve designed a proper intake manifold, and you understand the "black art" of carburetors.
After measuring both scoop height effectiveness and drag, the real test is the driver and his vision. It’s not likely that you’ll be afforded the luxury of the 11" scoop, but the height we can work with when combined with bottom end clearance and crank C/L will allow a maximum "package height" of 24.5 (the maximum distance from the top of the intake manifold to the crank C/L) and this number still provides the proper room on top for the carbs. and scoop relationship to work well. Next the decision to run lower rpm is examined, and a torque number is determined to be the minimum to push our combination of frictional losses, drag coefficient and frontal area to speeds of 200+. That torque value at 9,000 rpm is 792 ftlbs., or 1,360 hp. for those impressed by that #. As we’re going to run less rpm, our rpm range will be slightly wider by necessity, so the math shows that the lowest rpm we’ll encounter is 7,800. The torque necessary to overcome the shift is 852 ft.lbs. which for those who care is 1,265hp. Nasty numbers for a Detroit relic!
The bore / stroke combination for this rev. range will use a slightly longer stroke, and smaller bore, at first the first "look". This is where the package height starts affecting everything we do. Since I attempt to run a rod length to stroke ratio of 1.75 - 1, calculating the rod length combined with the compression height(distance from pin C/L to deck), and (.5 x stroke) will yield our block deck height once we settle on the stroke. The combination using 4.6" bore, and 3.75" stroke, with a 6.5" rod, and a piston with a 1.265" compression height will yield a deck height of 9.635"and the rod length to stroke ratio will be 1.73 -1 which is "good". As we’re looking for a really "fat" torque curve, we’re going to build another combination for this engine. Number (2) engine , combined will use a 3.80" stroke and a 4.57" bore combined with a 6.57" rod will again provide a rod to stroke ratio of 1.73 -1. In order to place each engines internals under the same deck height, the compression height on #2 will be 1.165" which is "doable".
So "derived from experience" I’m going to build a 498 cid. engine with two different strokes and bores. The two groups of 4 cyls. displacements will be within .01 ci. coming in at 62.29 ci. and 62.28 respectively. The geometry is 1.73 -1 on each, and the ring package will be placed in the same position despite the different compression heights. Both bores are large enough where the flow rates and swirl characteristics will be very nearly the same.
The bottom end will use a cast iron block which has been "seasoned" and stress relieved. All finish machine work will be done in house including main bearing bores, cam bore (which we’ll raise) lifter bore location, all oil passages, cylinder bores, bolt holes, and all finish interior and exterior machining. The exterior milling is for weight removal and once the mechanical operation is complete, the exterior is chemically treated to remove more material and to dispose of any tool marks which could serve as origins for stress cracks. The cam is moved up to shorten the pushrod length, and the lifter bores locations are also dictated by valve and rocker arm position. We want everything as direct as possible and no push rod angle relative to both lifter body and rocker arm. The camshaft will rotate in needle bearing housings pressed in the block, do oil isn’t a great issue. The lifters are of the roller variety, and their bodies are about twice the diameter of "stock" lifters. The lifter body is made of beryllium for weight purposes, and lifter rotation is prevented by a pin similar to a key which will ride in a slot in the bore.
The crankshaft is a non twist forging from a Japanese vendor. After rough machining it’s checked for internal and external flaws, and then the finish machining begins. We’ll use the big block Chevy main bearing size, and the rod journal will use a bearing which will have adequate width, but a specific diameter calculated to reduce bearing speed, and resultant losses. Once all the bearing surfaces are "close" the final shape of the counter weights are determined and they are machined to an aerodynamic configuration.
We elect to use titanium rods which will allow us to run tighter clearances than their aluminum counter parts. (piston to head @ .30" vs. .90") These rods cost three times more but they’re good for three times the life, or about 45 runs. The entire rotating assembly is internally balanced and ready for assembly.
The oiling system consists of a five stage dry sump pump, with external oil storage tank. The oil pan is designed with a large "kick out" on the rt. side where the oil pick-ups are located. We also fit a scraper to the shape of the rotating assy. That has about .050" clearance to shear oil from the crank / rods. The bottom of the oil pan is configured and coated to prevent oil from bouncing back up into the rotating assembly. Our pump not only removes oil but one stage is used to create a "vacuum" in the crankcase, the rest is used to pump oil back in. The engine will never have more than 1 quart of oil in it at any given time or rpm.
On top, things get interesting, because now we’re dealing with the piston domes, heads, valve events and manifolding.
During the selection of "brand" we automatically were zoned in on about 4 different heads which are legal. As we’ve had experience with these pieces before and we do not want to do a lot of welding, we’ll make runner, chamber, and water jacket cores which will fit the OE core box and provide heads with adequate material everywhere we need it. The "desirable flow #’s at all valve lifts are calculated, and the port angles and shapes are modeled to verify the characteristics. This is a critical stage as we must also design the intake manifold at the same time, and, remember, it all has to fit in the "package height" which was dictated at the beginning. We begin with the plenum top and overall volume. The volume is dictated by displacement, rpm and rod / stroke ratio, so it’s "fixed", however, a carbureted engine must have the correct runner length, angle, volume, and opening relationship relative to the carbs. throttle bores or the "vacuum signal" will not permit proper carb. function, and the engines ability to accelerate and recover from an instant drop in rpm (on shifts) will be non existent, regardless of torque and power #’s. Now we’ll begin at the top and work down. Once we have the carbs positioned correctly for easy runner entry access, we’ll determine if the intake port previously modeled can coexist with a runner of this angle. If so, we’re OK, but typically we need to go back and redesign the port to compromise its shape to allow it to be properly manifolded. We’ll go round and round and finally settle on a best of the lot. Keep in mind that this is a push rod engine, so the ports must be placed between the push rods, and remember the push rods must be as straight as possible, or valvetrain life is going to be non existent.
Things are now going to become a little more challenging. Remember that the engine has two separate sets of cylinders that do not share the same bore and stroke dimensions. The manifold runners also need to be different lengths and volumes to accomplish what we’re asking. So now we need four runners of one configuration, and four of another. We determine that the short stroke port / runner volume should be 1038.2 cc’s, and the long stroke will want 1128.6 cc’s. These volumes should maximize the cylinder’s out put in the desired rpm range.
We now design the combustion chamber and piston dome. We have various models which are known quantities, so we’ll select one that at first pass appears to want to work with bore size, port location (push rod location),etc. The primary goal is to create a really fast burn in a "big" bore, and to also allow great tuning tolerance. The other requirement is that we achieve maximum cylinder pressure well past TDC, at 20 to 30 degrees if possible. To achieve all this we’ll make sure that at TDC there’s no secondary pocketing. The chamber and piston shapes will accelerate the swirl initiated by the intake port’s flow bias, and then concentrate the swirling mixture into a small "sweet spot" near the exhaust valve, and to make things lively we’ll also put the plug in the "sweet spot" as well. The chamber shape now dictates valve location and size, however, if the valve size isn’t sufficient, we must return and look at remodeling the chamber.
I believe we have a chamber and piston dome that’ll work, and based on rpm and relative efficiency, the flow rates are determined for all areas under the lift curve, which is a function of how quickly I can open the intake valve and not exceed piston velocity for the first 12 - 14 degrees of crank rotation. Each intake valve will have it’s own lobe configuration and timing. This may seem acceptable for the differently configured cylinders, but it’s necessary to make changes on every cylinder as well to achieve our power range goals. The valve sizes will be 2.48" for the intakes and 1.87" for the exhausts. These aren’t as tight a fit in the bores as you may think, as there’ll be some compound angles involved so the valves will open to the center of the cylinder, and move away from the cylinder wall as they open. The "exit" area size for the exhaust port will be 3.04 sq. in., and the area inside the port will be as small as 2.0 sq.in., and I can’t tell you where relative to the port length as it’s highly proprietary.
The intake port cross sectional area will range from 4.84 sq.in. to a minimum of 3.23 sq. in., and again, I can’t say where these areas are located in the port. In the intake plenum chamber the runner entry will be 6.7512 sq, in. with a 1.46" radius around the runner entry.
There’s also a "black art" that I’ll mention regarding state of the art cylinder heads, and that’s steam manifolding. The design and construction are a pure art form, and if it’s not done correctly, you won’t be in the hunt. That’s all on that subject.
The flow rare for the intakes is 618 cfm @ .800" lift and 675cfm @ .800" for the exhaust ports. These values are only for comparison and are at a pressure drop of 28" H2O. We develop and flow components at "different" pressures than the "norm". The exhaust port will reach 85% of the above # by .400" lift. These lift #’s are not necessarily indicative of the lift #’s currently run in prostock as some cams yield over 1.0" net lift. I will say this, however, I feel the same way about valve lift as I do about rpm.
The material for intake valves is beryllium and the exhaust valves are titanium. The retainers are titanium, as are the valve springs which need replacement every 12 runs, and you already know the price. The shaft mount rockers are also beryllium, and the pushrods are composites.
The use of ceramic and non friction coatings is extensive on piston domes, skirts, combustion chambers, and valves. We also coat bearings, oil pump components, etc. with moly coatings.
The headers are 2.5" diameter primary tubes @ 25" long, the collectors are 4.5" dia. And 13" long after the merge is complete, and all exhaust components are ceramic coated


The carburetors are Holley 4500 series units that flow @ 1300 cfm. each. Fuel systems require a minimum 1" diameter line from the fuel cell to the carb logs, while maintaining constant pressure of 8 psi. The vent on the fuel cell must be at least .625" or you’ll starve the engine. The ignition is crank triggered MSD with multi- step rev limiters, and timing alterations.
Typical procedure is to break in the rings with a mild cam, "break-in "heads and induction components. Once complete and we see negative crankcase pressure, meaning the rings have seated, it’s time to see if the combination works.
After installing all the "good" stuff, we’ll do a quick run in to re-torque and reset the valves. The next hour is spent installing pressure transducers for each cylinder, exhaust temp. sensors, and chemical analysis probes in each primary tube, correct air ducting and inlet temp. and volume measuring instruments, and also the fuel flow and all normally used sensors. We have the ability to not only adjust carb mixture remotely, but we also can re-map the ignition timing. Dyno’s have come a long way during the last twenty years, and . Dyno’s have come a long way during the last twenty years, and we’ve attempted to stay on top of the game, as the more you can simulate in the dyno cell, the less track time is required. Aside from being able to program a drag strip, oval, or road course into the system, we also simulate G’s by mechanically tilting the entire dyno / engine combination. For lateral G. the engine is all we tilt.
We now have the ability to not only map cylinder pressure from each cylinder vs. crank angle at all rpm., analyze the combustion efficiency by looking at the gasses that are spent, but we can also read the torque of each cylinder during the tests, and when the test is complete, we can down load, and determine where we need to fine tune.
This particular engine produced slightly better numbers than we "dictated" on it’s first "run". We made some slight mixture and timing changes and found less torque but quicker acceleration. After another hard look at all the numbers, I felt that we needed to alter the cam’s lobes slightly to increase the torque overlap from the different 4 cly. Engines, but also the individual cylinders needed slightly more "spread".
Once installed and all was re-set, we realized considerably more torque than earlier(which was a surprise), but the response time was the best we’ve seen from a Pro engine. Not only did it immediately get up after the simulated shifts, but the transient response (acceleration time from point A to B) was incredible. We also observed that the best timing was 12 degrees BTDC, and the exhaust temps. were down around that 800 degree # we like. Chemically, it was also very clean so we were not only purging the area above the top ring, but also there was little that wasn’t burned. The optimum BFSC’s were all under .340 which is very good for any engine, and typical of what we’ve been seeing for the last 15 years or so. Most drag racers don’t care much about mileage though.
We did a few tests to determine where the engine became semi-resonant, and although it’s a large engine, the torque extended down to some lower rpm ranges than expected, and the combustion efficiency was surprisingly good in the lower rpm ranges. Although I’ve never been to build two engines that will produce identical power, it was decided that since almost everything was digitized, we’d build a clone to play with while freshening up the customer’s engine. Or is that engine(s)???
And for all those who want compression ratio numbers, the "combustion space" volume for each cylinder equaled 43.4 cc including the head gasket. If you calculate it, please don’t tell me what it is, as it might worry me and my customers! The gas requirement was 113 octane, and the individual cylinder ignition timing was as low as 11, and as high as 13 degrees….different engines, every cylinder
this is from t.o.o.and is his not mine just wanted to share what goes into a pro stock engine,instead of just giving hints
 
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