In this chapter, I go into more depth on power-producing factors that may have only been touched on briefly in previous books, and you may not have seen the piston tech I provide.
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You could easily buy pistons that needed to be modified for clearance somewhere on the crown. The biggest difference you are likely to find between one piston and another is in the plug positioning. The piston’s spark trough, or flame ditch, can be up to 1/2 inch off from where the plug is. This needs to be checked, and in some cases redone so the flame path has a clear run to the rest of the chamber. Make sure no part of thepiston crown inhibits the flame front passage; cutting the spark ditch is always a move in the right direction. However, the passage of the flame through the charge is complex and understanding what goes on takes a lot of testing to find what is needed.
An area where much work is currently being done is in the design of the cylinder heads’ quench pad and its interaction with the piston crown. For years, it was assumed that the quench pad should be just a flat surface that the piston closely approaches. Now, this is proving to be not the case. But it is not only the pistons’ quench area that interacts with the head; some less than obvious aspects of the piston design also affect the airflow.
Due to the fact that a performance Chevy big-block piston needs to generate a high compression and accommodate a high valve lift, especially on the intake, it presents some issues related to the breathing ability of the cylinder as a whole.
The issue with the valve cutouts is easy to identify, but there is another flow-inhibiting issue that is rarely appreciated. What you read in the following paragraphs may be your first introduction to flow-inhibiting factors for pistons.
The edge of the intake valve pocket needs attention and it’s the easiest area to take care of. As can be seen in Figure 2.4 the air entering the cylinder around the short-side turn during the overlap period runs into the wall of the valve cutout. This needs to be rectified as shown in Figure 2.5. This “piston porting” exercise is easy enough to appreciate if you spend a little time studying these figures. And it definitely delivers performance benefits.
However, one aspect that is mostly peculiar to 24-degree heads is far from intuitive. A pressure/flow distribution plot (Figure 2.6) of the intake port reveals that the busiest exit area around the valve is not, as is normally the case with most other two valve heads, toward the cylinder’s center. Measuring seat exit velocities with a vented valve shows that the busiest section of the port is often toward the shrouded side. Once this becomes known it is easier to see why, when the block is chamfered in this area, the response is a sizable amount of extra output. The block chamfering, however, starts to aid fl ow after the valve is around 0.150 inch and more off the seat.
Less obvious is that the fl ow pattern on the cylinder wall side of the port is spiraling past the edge of the bore shrouded intake valve. In effect, air is corkscrewing past the edge of the intake valve at about the 10 o’clock position, and during the overlap, the dome of a high-compression piston can block this flow. This suggests that you not only need to cut the top of the bore as discussed in Chapter 1, Displacement Decisions, but you should also find out if the piston dome can have any negative influence on the flow into the cylinder other than the effects of valve shrouding from the aspects indicated in Figure 2.3. From Figure 2.6, you can see that the flow into the cylinder does not follow a pattern that is by any means intuitive.
This flow test and the port probing just prior to the intake valve show an important flow pattern. As unlikely as it may seem, the flow corkscrews off the edge of the valve on the cylinder wall side of the port and then proceeds over the edge of the intake valve and into the cylinder. At least that is the way it would go if there were no obstructions. Because this flow pattern is generally unknown, piston domes rarely have a form that makes allowance for it. Depending on the height of the dome there is a potential 10 to 15 hp to be had by some subtle and some less than subtle reshaping.
The best piston crown shape to have is a flat one or one with a shallow dish in it. Unfortunately that usually results in a really undesirably low compression ratio unless the short-block has a lot of cubic inches. The first move is to address the edge of the piston’s intake valve pocket as per Figure 2.5, which shows the previously mentioned piston mod. From here on out the valve shrouding reduction moves are a little more subtle.
The principle job of the top compression ring is to seal against the pressures experienced above it. To do this, it must have some radial load pressing it outward onto the cylinder walls. A relatively high-compressive preload can achieve this, but that means excess frictional losses on the induction or exhaust stroke. You need to increase the rings’ radial cylinder wall loading as the cylinder pressure increases. Gas porting the top ring groove is an attempt to do just that.
If maximum output for a given piston is the goal, gas porting is the way to go and is offered by most piston manufacturers. For drag racing, using a vertical style of gas port through the piston crown is the preferred method. However, in time, these can clog up, so for use other than drag racing, a horizontal gas port in the top ring groove is preferred.
Thinner rings are better than thick ones due to reduced friction; Total Seal rings are about as good as it gets because they seal tight. This is not just my opinion but the result of a lot of tests with rings at various gaps all the way down to the zero gap given with a Total Seal ring. Like the bores the piston rings need to have a low-friction prep. The first move is to use a very fi ne stone to remove the sharp corners from the edges of the rings’ outside diameter (OD). Then polish the rings with a Scotch-Brite pad until they feel really slick.
Connecting rod failure is rarely less than catastrophic. Fortunately, factory rods are fairly stout pieces but carry a lot of excess balance-pad mass. With work, they can be lightened considerably. Also, you can install a set of ARP bolts. But subsequently, your local machine shop needs to cut the caps and resize the big-end bore.
When all of this is done, you will have about half the money into them that a set of good aftermarket rods cost. I use the word “good” here because I have seen failures with one or two brands of “off-shore”–sourced rods. Analysis has shown that the material spec was way off what it was supposed to be. If you stick with the rods I show here, you should be in about as good a shape as can be expected.
The rod I use most (because it has proven time after time to deal with the prolonged dyno sessions my mule engines go through) is Scat’s ICR6385 rod. In addition to being strong, it is typically lighter than a stock rod even though it is 1/4 inch longer between centers. Also, it is very affordable.
My gas dragster that has a touch of nitrous and produces a little more than 1,100 hp uses these rods. As of 2014, these rods are on their sixth season. In engines of up to 850 hp and 7,500 for a quarter-stroker, they have so far appeared bullet proof.
If you want to spend a little more and get rods with even more strength that are 100-percent machined, you can do so without breaking the bank. Options to consider include those from Callies, Crower, K1, Manley, and Scat. When it comes time to spend money on connecting rods, remember that longer is always better. Usually, when building a short-deck block, a 6.385-inch-long rod is the best choice if you are looking for maximum inches. If you are building a tall-deck block (10.2 to 11.1 inches), rods are available up to 6.8 inches long.
There is more to crank selection than simply deciding what stroke and rear seal style is needed. All the bigger stock-displacement factory cranks are externally balanced, so it was either inconvenient to accommodate enough counterweight mass within the confines of the crank-case or there was simply not room. To balance the crank required additional counterbalance mass incorporated into the crank dampener and flywheel/flexplate. Although this Band-Aid fix is passably okay for the rear of the crank, it is undesirable for the snout because it leads to unnecessary bending moments about the number-1 main journal. Unless cost considerations require it, absolutely do not go with an externally balanced crank and dampener system; it needs to be internally balanced.
Another aspect of crank design you may want to consider is whether or not to go with a design having center counterweights. The subject of center-counterbalanced weights may not have previously entered your thoughts. To understand what it’s about, see Figure 2.30.
Without center counterweights, a “couple” (a rocking effect caused by two forces) acts on the center main bearing; it is brought about because of the displacement of the two rod journals on either side of the main bearing. By fully counterweighting each throw, this couple is considerably reduced. But how important might this be in the grand scheme of things?
In terms of engineering finesse, a center-counterweighted crank is the way to go. It relieves the center main of some bearing loads and reduces the bending action caused by a lack of counterweight at this position. But it’s not an open-and-shut case. Having those extra counterweights usually means a slightly heavier crank, although not by as much as you might think. In a center-counterweighted crank, some of the mass for the center counterweights is taken from the next counterweights out. This means that, in part, some of the mass for the extra pair of weights is realized by the crank designer moving some of the mass from adjacent counterweights to the center counterweights.
Building a street or street/strip engine with an internally balanced crank means you have eliminated an out-of-balance mass at either end of the crank. That’s such a big step in the right direction that it makes the issue of using a center counterweight or not far less important. The center main is big enough to take the added loads of the couple around it, so reliability is not likely to be an issue, at the level likely to be seen for engines up to approximately 7,500 rpm and 900 hp.
If I’m building a cost-is-no-object engine and call the shots on a billet crank that needs to be as light as possible, I would go with the center counterweights for an endurance race engine. However, for anything short of Pro Stock or Pro Mod, the inherently lighter weight of a non-center-counterweighted crank is a good, although minor, option to consider.
The crank for one of my 598 builds (see Fig. 2.31) is a relatively high-dollar custom item, and I elected to go without center counterweights. If your application involves endurance and/or RPM at a relatively high level for extended periods, such as marine use, a center-counterweighted crank is your best option.
The inertial and gas pressure loads imparted to the crank at each rod station cause the crank to twist torsionally from end to end. When these forces are imparted at some multiple of the crank’s natural torsional frequency, they build up out of all proportion to the amount caused by the original excitation force.
A 4.75 stroke crank that was grossly under-dampened illustrates this condition. At about 4,000 rpm the crank starts to resonate and the rod journals vibrate back and forth during their rotation by almost 2 degrees. As a result, the position of the center of the rod journal on this seemingly very stiff crank is moving back and forth by an almost unbelievable 0.080 inch. This cyclic motion superimposed on the typical rotation of the crank plays havoc with the cam dynamics because the cam is coupled directly to the crank via the timing chain and gears.
In an extreme case, 50 to 75 hp is lost on a 750-hp engine, and a typical loss is 20 hp. Furthermore, the crank breaks at a fraction of its normal life.
All this tells you that an effective crank dampener not only greatly extends crank life but also allows the generation of additional power.
It is neither practical nor possible to rid the rotating assembly of all torsionals. But if those torsionals are reduced to some practical lower limit, you essentially eliminate any problems. Assuming the cam and valvetrain have inherently good mechanical dynamics, an effective dampener (compared to an indifferent one) can reduce the peak-to-peak torsional vibrations. On a 4.25-inch-stroke crank, this means a reduction from about 1.2 degrees to about 0.375 degree. Anything under about 0.4 degree is acceptable, but it is possible to make worthwhile reductions even on this figure.
Bear in mind that the longer the stroke, the lower the natural resonant torsional frequency. In this book, I typically deal with strokes between 4.25 and 4.75 inches. A really effective dampener can, with such stroke lengths, reduce torsionals that occur in the important 4,000 to 8,000 range to a peak-to-peak torsional vibration of as little as 0.2 degree. But to get that you must be very selective in your choice of a dampener.
Size and Weight
Engine builders have a strong tendency to focus unduly on dampener diameter and weight. The thinking here is that a lighter dampener, having less moment of inertia, allows the engine to accelerate faster so it must be better for performance. In 99 percent of instances, however, that proves to be not the case. The dampener attribute you should rank as number one, if optimizing performance on the track is the goal, is its ability to effectively dampen those unwanted torsionals. Everything else comes in far behind that requirement. Although there are going to be instances to the contrary, my experience is that a good 8-inch dampener marginally edges out a 7-inch dampener in terms of engine output by the odd couple of horsepower for an engine in the 800-hp range, even though it’s typically 2 to 21⁄2 pounds heavier (about 30 percent greater moment of inertia). My advice here is that for endurance applications go with a good 8-inch dampener.
The important weight/mass in a dampener is the inertial mass on the outside of the dampener. If you want to save weight and a little inertia, an aluminum hub is an option. However, my thoughts here are that the aluminum hub doesn’t take as many removals before the all-important fit on the crank becomes questionable. Remember, unless that fit is super tight the dampener does not work as it should.
Also, because I am on the topic of dampener retention, be sure to use an ARP crank bolt to hold the dampener in place: no exceptions here.
Dyno shops with torsional vibration test gear are virtually non- existent. That being the case, how do you test the effectiveness of a dampener? Fortunately that is very simple. The more effectively the torsionals are dampened, the more horsepower the engine makes. I have spent quite a few hours doing torsional measurement but at the end of the day the best dampener is the one that delivers the best output.
If you have made the decision to go with internal balance, you have addressed the number-one issue in the right manner. Also be aware that a little underbalance or overbalance does not affect how smoothly the engine runs. The biggest factor affecting just how smooth the engine runs is how close to uniform the weights of the rods and pistons are.
Written by David Vizard and Posted with Permission of CarTechBooks