Though not readily appreciated, understanding and optimizing an engine’s compression and compression ratio is a valuable tool toward maximizing performance.
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A four-stroke (or four-cycle) engine is so called because, in the process of producing power, the piston passes up and down the bore four times. These strokes or events are the induction, compression, power, and exhaust. As you may suppose, the effective function of all are important for producing a high-output engine. Of the four, the compression stroke has far less obvious but more reaching implications on an engine’s optimal spec and its subsequent success as a power producer.
Obviously, the principal idea of the compression stroke is to compress the intake charge as effectively as possible, and to do so with minimal leakage. We need to remember that as we continue, because there are two principal factors associated with the compression ratio. The first is the calculated ratio, which we refer to as the geometric or static ratio. The second and equally important factor is how effectively, and to what degree, the physical components of the engine compress the charge into the combustion space. In essence, it’s a measure of how effectively theoretical compression ratio is translated into real-world pre-combustion cylinder pressure. This is commonly known as the dynamic compression ratio and is influenced by such things as ring and valve seal and, to the greatest degree, valve opening/closing events.
Now you may well have heard the term compression ratio (CR) many times but may not know exactly what defines it and how it’s calculated. If so, you need to refer to sidebar “Compression Ratio Definition” on page 146.
More on the Strokes
It may look like we are treading a well-worn path here, but it’s worth taking a quick look at the four strokes, because each of the other three is intimately tied to the compression stroke. Every one of these strokes must accomplish its goal effectively for an engine to be able to produce a high output.
Let’s start with the intake stroke. The more efficiently the cylinder is filled on the induction stroke, the more RPM the engine can turn before it “runs out of breath.” The better the intake filling is, the higher the pressure achieved on the compression stroke. This, along with as high a compression ratio as the fuel permits, means significantly higher pressures on the power stroke.
On to the compression stroke. The higher the compression ratio, the higher the resultant combustion pressure. Not only that, but the charge burns faster, thus necessitating less advance for an optimal burn event. A higher CR also reduces the amount of residual exhaust remaining in the chamber at the beginning of the intake stroke. This reduces unwanted intake dilution by the exhaust. These are the most obvious power-enhancing factors, but they are not the biggest influencing factors. There are other less obvious but more influential implications that I cover later, when we look at the CR and compression factors in detail.
Next is the power stroke. Every bit of power the engine develops is made on this stroke. We need to make sure everything that happens before, during, and after this stroke either enhances it or at the very least has minimal negative impact on it. That means sealing the cylinder in the first place, making sure it does not leak throughout the power stroke, and ensuring that its sealing ability is not at the expense of high ring-to-cylinder-wall friction.
Finally, the exhaust stroke. Here, we need to make sure that cylinder emptying is done without undue pumping losses. Any pressure remaining in the cylinder while the piston is on the way up the bore is negative power. As far as exhaust stroke efficiency is concerned, having a higher CR can lead to significantly reduced exhaust pumping losses.
Thermodynamics Made Easy
It is easy to understand that increasing the CR raises cylinder pressures, thus causing torque output throughout the RPM range to simply follow suit. What is less obvious is that the increase in output from the higher CR comes about largely due to an increase in thermal efficiency. Thermal efficiency is a measure of how effectively the engine converts the heat-generating potential of the fuel, when burned with an appropriate amount of air, into mechanical power.
To more clearly appreciate how the thermal efficiency is improved, we need to consider the opposite side of the CR coin. This is the expansion ratio (ER), which describes what occurs as the piston moves down the bore on the power stroke rather than what happens as it moves up on the compression stroke.
Take a look at Figure 12.2 and then let’s go through the characteristic difference (computed taking into account typical heat losses) between a high-compression cylinder and a low-compression cylinder. For a moment, imagine that both the 15:1 and the 2:1 cylinders start off at TDC with 1,000 psi. As the piston of each cylinder moves down the bore, the drop in pressure follows a distinctly different line. The 15:1 cylinder drops pressure much faster than its 2:1 counterpart because of its more rapid change in volume. It only has to go down the bore a short way for the original volume to have doubled, whereas the 2:1 cylinder must travel to the bottom of the bore to double its original volume.
At the bottom of the stroke, the 15:1 cylinder has dropped to about 25 psi above atmospheric pressure, whereas the 2:1 cylinder is still at some 260 psi. In simple terms, the high-compression cylinder, when the exhaust valve opens at BDC, is only dumping 2.5 percent of its original pressure while the 2:1 cylinder is dumping 26 percent!
Up to this point, we have assumed that both cylinders start with 1,000 psi. But the best that the 2:1 cylinder really generates is about 200 psi. That produces the lower curve (light blue line) in Figure 12.2. The 2:1 and
15:1 cylinders both draw in about the same amount of fuel and air. But we can see that the 15:1 cylinder has more area under the curve by an amount equal to the green shaded area. Adding the green shaded area under the curve amounts to nearly doubling the power output from the same amount of fuel and air. That means, from the same heating value of fuel, we have doubled the thermal efficiency, and in so doing extracted twice the power.
You can now see why a high-compression cylinder produces better power and fuel economy. It is not solely because the charge is squeezed harder and the resulting combustion pressure is increased, but also because the higher expansion ratio allows more energy to be extracted from the original high-pressure charge.
Simple Theoretical Power Gains
Figure 12.3 can be used to calculate the theoretical power gains seen from raising the CR, and Figure 12.4 saves you the effort of calculating those gains. This formula does not take into account the inevitable heat losses, and 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 14:1. From there on up, some heat is absorbed by chemical reactions brought about by the high temperatures and pressures generated. This heat is subsequently delivered back to the cycle, but it’s too late in the expansion event to serve any useful purpose.
Because of this, many books 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 the side benefits of ultra-high compression are taken advantage of, the situation takes a complete about-face.
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 (Figure 12.3) works to produce results far better than theorized. In other words, all the numbers in Figure 12.4 are on the low side. For instance, a mildly modified 9:1 350 small-block Chevy makes about 380 ft-lbs of torque. Based solely on our thermal efficiency formula, raising the compression to 12:1 should bump that figure to 397 ft-lbs.
In practice that number is usually exceeded and the bigger the cam, the bigger the gain. To understand how much more can be had, let’s look at the effect the cam has on the situation.
At lower RPM, the static CR is never realized because our thermal efficiency formula assumes 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 does not close the valve until about 50 degrees after BDC. Figure 12.5 shows the typical extent of piston motion back up the bore before the intake closes for three cams.
Because of the delayed intake closure, there is considerable piston motion up the bore from BDC before the intake actually closes. This, at low rpm, pushes some of the mixture 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 100-cc cylinder with a static CR of 10:1 may only trap 75 cc 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.
The following example shows just how much influence the delayed intake closure has on the dynamic CR. Let us take three different duration cams, all having a 108 lobe centerline angle (LCA), and all timed-in at 4 degrees advanced. Along with this, let’s say our static CR measures 12:1. With a 250-degree duration cam, the dynamic CR is in the mid to low 11s. For a cam of some 275-degrees duration, the dynamic CR drops to around the mid 10s.
Because of the piston, rod, and 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 when we get much past about 280 degrees, we had better have a decent dynamic CR. To give you an idea of the extent to which this occurs, our 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, maybe the dyno test results in Figure 12.6 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 2-liter version of this engine, because of its geometry, reacts just 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 265-degree cam produced (the gray curves of Figure 12.6) 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 just shy of 140 hp. This cam was then substituted for a 285-degree cam.
On the same 9:1 CR (blue curves of Figure 12.6) this bigger cam dropped 38 ft-lbs 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 ft-lbs 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 Figure 12.6. As you can see, this increase in compression regained almost all the low-speed torque that was lost. On top of this the big-cam/high-compression combo produced an increase of 15 ftlbs and 33 hp. Projecting that result to equate to a 350-ci engine, the numbers look more like 40-plus ft-lbs extra and 95 hp. Are these numbers realistic? Sure they are. I have seen a 100-plus-hp increase from a 355-ci small-block Chevy with 25 degrees more cam duration, 0.100 more lift, and 2 points more compression.
When we go back to the basics, the big increases seen with a combo of more compression and cam are easier to understand. If you check the numbers in Figure 12.4, you see that the biggest gains from a compression increase happen when moving from a low compression to a higher one. Going from 8:1 to 10:1 is worth a theoretical 3.7 percent while 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.
By now, some of you are wondering whether the engine you have just built has enough compression for the cam you chose. Assuming your engine has good ring and valve seal, a simple way to determine whether 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 I discuss how to achieve that on page 154. If the cylinders are sealing 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.
How effective a compression test may be, for determining 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. The best way to establish that is to do a leak-down test. This requires a leak-down tester and a source of compressed air at about 100 to 110 psi.
Just how much leak-down is acceptable is open to debate. With the rings and bore prep that 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 if your cylinders check out at 7 percent or less, you are okay. With such a cylinder, let the compression gauge go eight pumps and use that as a reading to determine your cam/compression compatibility. If the ring seal shows 10-percent leak-down, that’s borderline for a high-performance engine and compression readings are going to be artificially low. If the leakage is 15 percent or more, maybe you should consider new rings to be a performance-enhancing move, as much as a reconditioning one.
Intake to Exhaust-Valve Ratios
The controlling factors influencing the best intake/exhaust ratio for maximum output has been much debated for the past half-century. (Of course, this assumes all the available space for valves is used.) For the most part, it 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/ exhaust ratio could range from as little as 1:1 (for a low-CR supercharged engine) to as much as 1:0.6 (for a very-high-CR NA engine).
What is usually not appreciated 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 the exhaust valve opening can be made earlier, and it can be held open longer. This can be done for an improved high-RPM output without significantly impacting the engine’s low-speed output. The rule here is that the higher the compression ratio goes, the smaller the exhaust valve 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, the exhaust valve needs to be left on the seat until later in the power stroke to avoid unnecessarily dumping usable cylinder pressure. Because it has to open later, there is less time to dump the exhaust, especially in the blow-down phase, 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 10:1 to 13:1 range, but by the time we get to 16:1 plus, the optimum is to have the exhaust flow at about 65 percent of the intake.
Maximizing High-Ratio Results
By now, it is pretty clear that making the most of the potential that can be had from high compression is a goal worth pursuing. But as the ratios sought get higher, problems can begin to arise. Probably the most commonly seen of these is due to the final combustion-chamber shape achieved when all the stops have been pulled out.
The problem here is, as ratios much above about 10:1 are required, the only way to further minimize the volume after maximizing head milling is to have a raised-crown piston. To a point this is okay, but if the crown intrudes into the chamber too far it can, as previously mentioned, severely compromise the flame travel, resulting in a very ineffective combustion process. How much can be lost? Suffice it to say, I have seen 100 hp disappear because of a piston crown intruding about 1/8 inch too much.
The rule here is, unless you know what combination of chamber and crown form works or are prepared to do the necessary R&D, don’t go overboard on crown intrusion into the chamber. For typical small-block V-8s from Chevrolet, Chrysler, or Ford, a good guideline is to use no more than about 0.100to maybe 0.125inch crown height in your quest for a high CR.
If you are forced to stick with conventional heads patterned after the OE-style head, big-block Chevys can be something of a law unto themselves. Compared to a regular parallel-valve engine, the chamber is somewhat less than conventional. A big-block Chevy tolerates a substantially raised crown before the tradeoff starts to cancel out potential gains. The key is to make sure the raised section of the crown does not too closely shroud the spark plug.
If achieving the CR sought results in an overly intrusive crown, there is an alternative solution. Instead of trying to reduce the capacity of the combustion chamber, try increasing the capacity of the cylinder. Either a bore or stroke increase does this. For instance, if you were looking to achieve, say, 10.5:1 with a 454, it would take a maximum head milling job plus a piston intrusion approaching 1/2 inch. The head-milling job is going to mean a lot of possibly expensive manifold machining to re-align the ports. An easier and only minimally more expensive way is to install one of Scat’s cast-steel 4¼-inch stroker cranks.
This and a 0.100 overbore delivers 505 inches and also allows you to achieve a 10.5:1 ratio with a very acceptable crown height of about 0.150. The same kind of move can be beneficially applied to small-blocks. Using an inexpensive stroker crank in a 350 Chevy not only delivers extra cubes, but also allows a 10.5:1 CR to be achieved with flat-top pistons and regular un-milled 68-cc heads.
The quench clearance is the distance the deck of the piston is from the cylinder head face at TDC. Loose (wide) quench clearances can actually promote detonation. The worst to have for most conventional-style wedge-head V-8s is about 0.100 to 0.125 inch. Reducing this clearance (by block milling or using a taller piston) can actually stave off detonation by a substantial amount. How tight the quench can be made depends on how flexible the block and bottom-end assembly is and how much thermal expansion has to be allowed for. With good steel rods and crank, the net clearance can usually be taken down to 0.030. With a typical Fel-Pro gasket of 0.040-inch thickness, this means the pistons come out of the block by 0.010.
If quench is so good at suppressing detonation and allowing the use of higher CRs for more power and better mileage, why doesn’t the factory make it tight to start with? In a nutshell, the answer is emissions. Tight quench areas cause unburned hydrocarbon emissions to increase. High-compression ratios bring about a dramatic increase in oxides of nitrogen, which are the primary cause of smog. Should we worry about this for our street machines? No; some high-flow cats and a well-calibrated fuel delivery system keep emissions adequately in check.
Containing the Pressure
Having a high-compression ratio brings about greater demands on cylinder sealing. The higher the pressures involved, the more attention you must pay to details. The first part of the equation for sealing the cylinder is to make sure your machine shop hones the block correctly. This should involve using a deck plate to simulate the distortion brought about by the stresses of head bolt tightening.
Then make sure your machine shop is aware of the type of piston ring material being used, so they can apply an appropriate finish.
Next, give the bores a good rubdown with a new Scotch Brite–type pad and plenty of Gunk engine cleaner. After that, scrub (with a stiff brush) the bores with a strong liquid detergent and rinse them with hot water.
When you are sure they are clean and grit free, immediately spray the machined surfaces with WD-40 to prevent rust.
With the bores ready, let’s look at the rings that are to ride on them. With modern oils, ring wear is hardly the problem it used to be; so use the thinnest practical rings. Many older-style V-8 pistons are still in wide production. The majority of these pistons still have 5/64 compression rings. There is no good reason for using these wider rings. The 1/16or even 0.043-inch-wide rings are what you should go for.
Be aware that the wider the ring gaps are, the greater the loss of cylinder pressure and, consequently, power. Add to this an increase in blow-by into the crankcase. This contaminates the oil faster and necessitates more frequent oil changes. If you stick with conventional rings, gap them to the minimum recommended by the manufacturer. If you can afford them, go with Total Seal rings because they really do deliver near-100-percent sealing and, equally important, they maintain it over a substantially longer period than even the best regular-type rings.
I mentioned the term “gas porting” earlier, but I am not quite finished with it yet. Gas porting is a technique to back up the top ring with combustion chamber pressure, so that the ring is more firmly pressed against the bore. There are two types of gas ports: those that pass down through the crown of the piston (see Figure 12.22) and those that are located radially, intersecting the top surface of the top ring groove.
The radial-style gas ports are common for long-distance race engines. The current trend is to use radial gas ports because they seem to be as effective, but do not unduly accelerate ring and bore wear at TDC. With a good race blend or street synthetic oil, bore wear at TDC is not really an issue. I have just completed a 1,000mile endurance test with the new Joe Gibbs Racing race oil and the rings of the gas-ported JE pistons in my Cup Car engine wore less than 0.0003 off the surface. This amount of wear led to the ring gap getting bigger by only about 0.0010. An oil analysis at the 100and 1,000-mile point indicated that most of the wear took place in the first 100 miles. This indicates that the ring and oil combination could be good for as much as 10,000 race miles.
Written by David Vizard and Posted with Permission of CarTechBooks