Simple Theoretical Power Gains
The nearby formula (see Fig. 1) can be used to calculate the theoretical power gains seen from raising the CR and the chart will save you the effort of calculating those gains. This formula does not take into account the inevitable heat losses, so to allow for this the value of "K" is commonly reduced from 1.4 to 1.3. Using this value we find that, changing nothing else but the compression, output pretty much follows the trend dictated by the formula until about 15:1. From there on up, chemical reactions brought about by the high temperatures and pressures generated absorb heat and only deliver it back to the cycle so late in the expansion event as to serve no useful purpose. Because of this, many learned textbooks will tell you that trying to utilize CR past about 14:1 is a fruitless exercise. But this only applies if no other changes are made to the engine. If, as we shall now see, the side benefits of ultra-high compression are taken advantage of, the situation takes a complete about face.

Dynamic Compression
In the real world, we normally find that theoretical increases are not usually seen in practice because of losses which, to simplify already complex theory, we have ignored. For high-performance engines, part of what has been overlooked by the simple thermal efficiency equation works to produce results far better than theorized. In other words, all the figures in the chart (Fig. 2) are on the low side. For instance, a mildly modified 9:1 350 small-block Chevy would make about 380 lb-ft of torque. Based solely on our thermal efficiency formula, raising the compression to 12:1 should bump that figure to 397 lb-ft. In practice, that number is usually exceeded and the bigger the cam involved, the bigger the gain. To understand how more can be had, let's look at the effect the cam has on the situation. At lower rpm we find that the static CR is never realized because our thermal efficiency formula makes the assumption that the intake valve closes exactly at BDC prior to the start of the compression stroke. This does not happen in reality.

At low rpm, port velocity and pressure waves are too weak to produce any cylinder ramming. Couple this to the fact that even a short cam of some 250 degrees of off-the-seat timing will not close the valve till about 50 degrees after BDC. Fig. 3 shows the typical extent of piston motion back up the bore before the intake closes for three cams. Because of the delayed intake closure we find that during the period the piston moves up the bore from BDC until the valve closes, a significant amount of the induced air is, at low rpm, pushed back into the intake manifold. This means the volumetric efficiency (breathing efficiency) and thus the effective displacement of the cylinder is well below 100 percent. In other words, a 100cc cylinder with a static CR of 10:1 may only trap 75cc of air. This means the dynamic CR, at about 8.5:1, has dropped well below the static CR of 10:1. The bigger the cam, the more this effect comes into play.

An example here will show just how much influence the delayed intake closure has on the dynamic CR. Let us take three different duration cams, all having a 108-degree Lobe Centerline Angle (LCA) and all timed-in at 4 degrees advanced. Along with this let us say our static CR measures out at 12:1. With a 250-degree duration cam, the dynamic CR will be in the mid- to low-11s. For a cam of some 275-degrees duration, the dynamic CR will drop to around the mid 10s. Because of the piston/con rod crank geometry, the piston tends to move much more slowly around BDC. This works in our favor for shorter cams, but the piston quickly moves out of this sweet spot, so once we get much past about 280 degrees we had better have a decent dynamic CR. To give you an idea to what extent this occurs, we find that with our example a 300-degree race cam used with a static CR of 12:1 has a dynamic CR of only about 8.3:1. This snippet of info should bring home the importance of having sufficient CR for a big cam. If it doesn't, then maybe the dyno test results in Fig. 4 will. These are some tests I did with the 2-liter Ford Pinto series of cams I designed for Kent Cams in England some years ago. I realize that very few of you drive Pintos, but the two-liter version of this engine, because of its geometry, reacts about the same as a typical small-block Chevy, so the results do directly apply. From these results we see that with a 9:1 CR, a 260-degree cam produced (the gray curves of Fig. 4) some decent results from low rpm on up. As expected, it started to drop torque by the time 5,000 rpm was being approached, and power peaked out at just shy of 140 hp. This cam was then substituted for a 285-degree cam. On the same 9:1 CR (blue curves of Fig. 4), this bigger cam dropped 38 lb-ft of torque at 1,750 rpm. That amounts to a 32 percent reduction. The extra duration did not start to pay off until 3,750 rpm. From there on up the bigger cam paid off by delivering an increase in peak torque of 4 lb-ft and almost 26 hp. At this point the head was milled to bump the CR to almost 12:1. The results of this move are shown by the green curves in Fig. 4. As you can see this increase in compression recouped almost all the low-speed torque that was lost. On top of this the big cam/high compression combo produced an increase of 15 lb-ft and 33 hp. Stepping that result up to a 350-inch engine, the numbers look more like 40-plus lb-ft extra and 95 hp. So are these numbers realistic? Sure they are. I have seen well over 100hp increase from a 355-inch small-block Chevy with 25 degrees more cam duration, 100 thousandths more lift, and 2 points more compression.

The big increases seen with a combo of more compression and cam are easier to understand when we go back to the basics. If you check the numbers in the chart (Fig. 3) you will see that the biggest gains from a compression increase happen when moving up from a low compression to a higher one. Going from 8:1 to 10:1 is worth a theoretical 3.7 percent, whereas raising the compression the same two points from 11:1 to 13:1 is only worth 2.5 percent. This means the bigger the cam, the more responsive it is to an increase in CR, especially in the lower rpm range.

Compression Pressure
About now there are going to be some of you who are wondering whether the engine you have just built and installed has enough compression for the cam you chose. Assuming your engine has good ring and valve seal, a simple way to determine whether or not this is the case is to check cylinder compression pressures. With the ring package and bore prep procedure I use, my own engines are almost always near zero leakage and we will look how to achieve that later. If the cylinders are sealing up well, I look for 190 psi as a lower limit with preferably 200 psi as a target when using 93-octane fuel. For every octane number less than 93, the compression pressure needs to be about 5 psi less to avoid detonation under normal circumstances.

As good as a compression test may be, to determine whether or not the cam you are using is accompanied by adequate compression hinges, to a certain extent, on how well the rings and valves seal up. The best way to establish that is to do a leak-down test. This will require a leak-down tester and a source of compressed air at about 100-110 psi. Just how much leak-down is acceptable is open to debate. With the rings and bore prep I use, I expect no more than 1 percent and something close to zero is what I normally see. But the average street engine is rarely that good, so we will talk in terms of more practical numbers. If your cylinders check out at 7 percent or less, then you're okay. With such a cylinder, let the compression gauge go 8 pumps and use that as a reading to determine your cam/compression compatibility. If ring seal is such as to show 10 percent leak-down, then it's borderline for a high-performance engine and compression readings will be artificially low. If the leakage is 15 percent or more, then maybe you should consider new rings as a performance-enhancing move as much as a reconditioning one.

Intake-To-Exhaust-Valve Ratios
The controlling factors influencing the best intake-to-exhaust ratio for maximum output (and this does assume all the available space for valves is used) has been a much-debated subject that, for the most part, has left the reader little or no wiser. The often-touted 75-percent rule is usually accepted without further question. In reality, the value is far from fixed. The optimum intake-to-exhaust ratio could range from as little as 0.75:1 (for a low CR supercharged engine) to as much as 1:0.6 (for a very high-compression naturally-aspirated engine). What is usually not appreciated here is that the CR is, for the most part, the controlling factor. Because the high-compression cylinder delivers energy to the crank much earlier in the power stroke, there are implications we can take advantage of. The most obvious is that the exhaust valve opening can be made earlier and held open longer. This can be done for improved high-rpm output without significantly impacting the engine's low-speed output. The rule here then is that the higher the compression ratio goes, the smaller an exhaust valve is needed to get the job done. This in turn leaves more room for a larger intake.

When we are forced to use a lower compression, such as in the case of a supercharged engine, then the exhaust valve needs to be left on the seat until later in the power stroke so as not to unnecessarily dump usable cylinder pressure. Because it has to open later, there is less time to blow down the exhaust, so a larger valve must be used at the expense of the intake. That 75-percent exhaust-flow rule mentioned earlier works for engines in the 10 to 13:1 range, but by the time we get to 16:1-plus, the optimum is to have the exhaust flow about 65 percent of the intake.