Perfecting the Combustion Process for More Power – Part 11

I really wanted to title this chapter “In-Cylinder Turbulence and Combustion Dynamics,” but it wouldn’t quite fit. Nonetheless, it is where we are going here, because cramming a cylinder full of air is just one aspect of making power. Exactly how it is filled, and what we do with it once filled, still has a great deal to do with the production of power.


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Defining Combustion

Before even getting into our discussion of “in-cylinder mixture dynamics,” I want to make sure we are on the same page when it comes to the term combustion, as it relates to the internal combustion engine. I want everybody to understand this, so I offer no apologies for oversimplification. First let’s look at the overall, grand scheme of things. Refer back to Figure 1.1 in Chapter 1. It is the summation of everything you need to know about the production of power from an engine.

In the “Long-Block Assembly” area of that illustration, there are more factors to take care of in the “Optimize Cylinder Head Airflow” box than in any other box. To do this we have to take care of the six issues quoted within that box. That is our goal here. And to start with, let’s take a very basic look at what goes on in the cylinder.

Burn: Yes, Explode: No

First, the charge in the cylinder does not explode when the plug fires. It burns and the rate at which it burns in no way resembles an explosion. In a NASCAR Cup Car engine at peak RPM, the charge burns at approximately 150 mph across a 4-inch bore. That’s slower than the car is usually racing! An explosion is something that ignites at more than 2,000 mph—and dynamite burns a whole lot faster than that. At a 700rpm idle, the flame speed barely makes 10 mph! This is why we have distributors with ignition advance and advance curves.


Fig. 11.2. The orange arrow shows the intake port pressure sensor. The green arrow is the equivalent for the exhaust. The blue arrow is the cylinder-pressure measuring sensor that is in the form of a spark plug.

Fig. 11.2. The orange arrow shows the intake port pressure sensor. The green arrow is the equivalent for the exhaust. The blue arrow is the cylinder-pressure measuring sensor that is in the form of a spark plug.

“It’s an explosion that pushes the piston down the bore.” “It’s the burning fuel that pushes the piston down the bore, not the air. The air is there just to support combustion.” “It’s the flame that pushes the piston down the bore.” All of these statements are erroneous and demonstrate the fact that a working understanding of what is involved is not common. Before we can go on to look at, and hopefully optimize, the combustion process, we need to understand some basic principles. The easiest way to do that is to start with an external combustion engine, i.e., a steam engine. With this engine, combustion takes place outside, or external to, the working cylinder.

Two factors make a steam engine work: 1.) combustion of a fuel to generate heat and 2.) a working medium. The fuel for heat can be oil, coal, or wood, but in every case, the working medium for a steam engine is steam generated from water. Heating water creates steam at high pressure. This high pressure is then applied to a piston/crank mechanism and turned into motive power. The key factors to note here are: the generation of heat and a working medium.

Let’s now consider the internal combustion engine. For “internal combustion,” we could substitute “internally generated heat source.” Let’s do just that. Imagine a cylinder full of air that has been compressed by the piston going to the top of the bore. Instead of burning gasoline to provide the heat, we substitute a piece of thick wire and put a high voltage and several thousand amps through it. Assuming it is suitably sized, the wire acts like a fuse and instantly vaporizes. The heat it produces expands the air (such as what a typical-size V-8 cylinder could do) on this power stroke, and generates about 50 hp.

MuscleCarB

The interesting thing is that most of what is released from this cylinder is not exhaust but hot air and atomized particles of metal. From this example, we can see that heating the air places the pressure on the piston, and it’s the expansion of this hot air that pushes the piston down the bore.

Unlike with a steam engine, which uses air only to provide the oxygen for combustion, the air taken into an internal combustion engine has to serve two purposes: It has to supply the oxygen for combustion and it also has to be the working medium. My point here is that it is the heat of combustion expanding the air that supplies the power. If we could, at the moment after combustion, somehow suck all the heat out of the air in the cylinder, the pressure would immediately drop back to whatever the compression pressure was at that point in the stroke.

So, what we can say for sure at this point is: We need to generate the most heat possible by fully utilizing the heating value of the fuel, and also make sure that, after producing heat, we don’t aimlessly squander it. All this falls under two headings: combustion efficiency and thermal efficiency. Let’s define these two parameters so we all remain on the same page.

Let’s tackle combustion efficiency first. In practice, it’s a complex process that involves not only how much of the fuel is burned compared with what should have burned, but also how fast it was burned. Thermal efficiency is the amount of power developed at the flywheel divided by the amount of power that could have been developed had all the heat energy of the fuel been turned into mechanical work. In this second context it needs to be stated that the compression ratio, as discussed in Chapter 10, is the principal influencing factor toward high thermal efficiency. 

Combustion Efficiency

I rarely have access to the sophisticated and diabolically expensive equipment that the big auto makers use to do combustion research. The test gear I have been able to beg, borrow, or steal falls more into the super equipped hot rodder’s category. That being the case, the plan here is to bring to bear what is now my 50-plus years of race engine building and development experience. This means looking at the things we can, as hot rodders, practically apply to make our engines better. There won’t be too much SAE-type hardcore (hard to understand and nearly always impossible to apply) stuff, unless we can physically apply the info gleaned from such sources.  


Fig. 11.3. A currently well developed chamber for the A-Series engine from Swiftune. It has changed dramatically from those early heavily shrouded designs.

Fig. 11.3. A currently well developed chamber for the A-Series engine from Swiftune. It has changed dramatically from those early heavily shrouded designs.

 
Fig. 11.4. In-cylinder sensors were historically very expensive but TFX has managed to drastically reduce cost while increasing accuracy and reliability.

Fig. 11.4. In-cylinder sensors were historically very expensive but TFX has managed to drastically reduce cost while increasing accuracy and reliability.

 
Fig. 11.5. The data acquisition box for the TFX system is small and can be used in the dyno cell and as part of an on-board data acquisition system.

Fig. 11.5. The data acquisition box for the TFX system is small and can be used in the dyno cell and as part of an on-board data acquisition system.

A-Series Heads  

You recall that my first experiences modifying cylinder heads were with the Weslake-designed A-Series heads for Mini Coopers and the like. As poor flowing as these heads were, though, I learned a valuable lesson about nottaking things for granted.  

Throughout the 1960s the hot ticket for induction on the A-Series engine was a big side-draft 2-barrel Weber carb. These things really snorted (and that’s not a figure of speech). About the late 1960s, the Dell’Orto side-draft carb, which is a Weber lookalike design, began to filter into the UK. The innards looked a lot more sophisticated than its Weber counterpart, and for a given venturi size, these big Dell’Ortos flowed more air. I reasoned that, since it bolted right in the same place as a Weber, a back to back dyno test would be about as simple as tests come.

However, before I could get on Tecalamit’s dyno in Plymouth, England, I happened to talk to ace Mini racer Richard Longman. Just a week or two before, Richard had used the Tecalamite dyno to test fuel injection and Dell’Orto carburetion versus the Weber on a race engine similar to the one I was going to test. He reported that, on his engine, the Dell’Orto lost 8 of the engine’s original 128 hp, and when the change to fuel injection was made, 10 disappeared. Frustrated over why, Richard put the injector nozzles into the Weber carb bodies. Guess what? That engine still lost 10 hp compared to using the carb’s own fuel-delivery system.

Although I didn’t know why, I thought that the reduced output with the carb/fuel-injection change might be a peculiarity of a Long man spec engine, so I went ahead and did my own dyno tests. The results mirrored the Longman tests. I now had a burning question that defied an immediate answer: Why did delivering a more finely atomized fuel spray cause the engine to drop so much power? The Weber carb was less effective at atomizing the fuel. In fact, fuel came out of the auxiliary venturi (booster venturi) in globs, rather than anything that resembled a spray. Yet that A-Series engine loved it; and to this day, I have never figured out why globs work and even a moderately good spray does not.

The Chrysler Avenger Saga

In 1970, Chrysler UK unveiled an interesting car, the Avenger. And somehow or other, I got involved with the factory, doing development work as an outside contractor. The Avenger was, from the ground up, basically an all-new machine, and came in two and four-door forms. John Clarke designed the engine and, fortunately for me, he was one of the numerous genius minds I have managed to surround myself with over the years. John had designed an engine intended to be as environmentally friendly as possible, short of having a catalytic converter. The combustion chamber on this engine was, for a modern engine at least, unconventional. It was formed by the piston stopping about 1/4 inch short of the top of the bore. The head itself had no chamber; it was flat. It was, in fact, a quenchless chamber.


Fig. 11.6. Chrysler’s Avenger engine was somewhat unorthodox; the combustion chamber was formed by a piston stopping about 1/4 inch from the deck of the block, topped off with a flat chamber-less head. The intake port was a sound design and could be ported to give very good flow.

Fig. 11.6. Chrysler’s Avenger engine was somewhat unorthodox; the combustion chamber was formed by a piston stopping about 1/4 inch from the deck of the block, topped off with a flat chamber-less head. The intake port was a sound design and could be ported to give very good flow.

Fig. 11.7. I so often have people tell me what I can’t do. In this instance it was win a drag race championship with a Chrysler Avenger. I was also told sedans don’t have what it takes to do flame burnouts!

Fig. 11.7. I so often have people tell me what I can’t do. In this instance it was win a drag race championship with a Chrysler Avenger. I was also told sedans don’t have what it takes to do flame burnouts!

Fig. 11.8. A spectator took this snapshot from the stands with a long lens. After being boxed in at the start at P4 (there were two 350 Z28s and a 3.4-liter injected BMW in front of my 1,600-cc Avenger) I was relegated to last into turn-one. This is on lap five and I have just taken the class lead from the previous year’s champ, Bill Sydenham. Having a better understanding of the combustion dynamics gave me a clear 30-hp advantage.

Fig. 11.8. A spectator took this snapshot from the stands with a long lens. After being boxed in at the start at P4 (there were two 350 Z28s and a 3.4-liter injected BMW in front of my 1,600-cc Avenger) I was relegated to last into turn-one. This is on lap five and I have just taken the class lead from the previous year’s champ, Bill Sydenham. Having a better understanding of the combustion dynamics gave me a clear 30-hp advantage.

The intake ports on this head, which were on the same side as the exhaust ports, were very strong on flow. I was supplied one of these cars for my research. It was a two-door GT with 7 miles on it. Later that year, Chrysler brought out a hopped-up version of this car, equipped with twin side-draft Webers and a ported head. This, together with some suspension upgrades and styling mods, was introduced as the Avenger Tiger. It was destined to go head to head with Ford’s Lotus Cortina. Although it matched the Lotus Cortina for handling, braking, and cornering, it lacked the power to out drag it.

Anyway, back to the Avenger GT that I had to work with. To get a good part-throttle burn and clean exhaust from the intake charge, the twin 1½-inch Strom-bergs were heated. This was achieved by having the intake manifold bolted to the exhaust manifold. Between the two was a 1/16-inch-thick plate with a hole in it. Through this hole the exhaust flame physically played onto the underside of the intake, forming a very hot spot. This, at part throttle, was more than sufficiently hot for vaporizing all of the fuel at any sane street or highway driving speeds.

Okay, this might sound like stock boring stuff, but now we come to the point of the matter. My first discovery was that if the hot spot was semi-eliminated, by replacing the 1/16-inch-thick plate with a hole in it by similar plates with no holes, the power dropped from 78 rear-wheel hp to 74, even though the charge temperature dropped by some 40 degrees. With a quenchless chamber, I had vaguely suspected this might be the case.

This in turn suggested that this type of chamber needed to have a fair amount of vaporized fuel and the rest delivered in really well atomized form. How I figured out the ideal setup for the intake and optimal output for the engine is what I next explain.

This was the year (1972) I had planned on drag racing the car at Santa Pod Raceway, England, in a class that can best be described as “improved street stock.” This class allowed the head to be ported. Any intake and carb combination that was available off the dealer’s shelf could also be used. A race exhaust system was permitted. The cam, valve-train, and everything else below the block deck face had to remain stock.

The factory offered a twin side-draft 40 DCOE Weber kit for this car, which Chrysler’s race team manager, the late Des O’Dell, gave to me to test. Knowing that this engine liked fine fuel atomization and that the Dell’Orto DHLA carb delivered much finer fuel atomization, I tried a set, along with the Webers, on the chassis dyno. The difference between the finer fuel atomization of the Dell’Ortos and the more blob-like delivery of the Webers produced a night-and-day difference on the dyno.

I  wrote a story about the test involving these two brands of carb and Weber’s UK boss apparently went ballistic—told people I did not know what I was talking about, the tests were rigged, etc. He sounded one step shy of putting out a hit contract on me!

The first drag race I did with the Avenger was somewhat amusing. I turned up at the track with this car, viewed by 99 percent of racers at the time as a genuine grocery getter, and the jokes were rampant. Yep, I was the guy to laugh at, before the racing started. After easily putting everyone on the trailer, there weren’t even any smiles as far as I could see. Even 1,600-cc four-valve-per-cylinder Cosworth BDA Escorts went down to my two-valve 1,500-cc pushrod, all-iron engine Avenger.

I continued to develop this engine and, by the end of the year, had it running really well by paying close attention to mixture and ignition properties as well as overall airflow. The useful power band was from 400 to 8,000 rpm (that is not a misprint). This engine had punch everywhere between those two numbers, and it had awesome torque even by today’s standards. It also put any VVT Honda to shame for shear drivability.

Clive Richardson, of the UK’s Motoring News, conducted a road test and it showed this car capable of 0 to 60 mph in just less than 6 seconds and 0 to 100 mph in 17 seconds. Another test done by Motor Magazine at the Lindley proving ground just outside Coventry, England, showed how strong the Avenger’s super-wide power band was. When compared to the figures delivered by a Ford Cosworth BDA Escort, the “through the gears” acceleration was significantly better (more than 1 second faster to 60, and 3 seconds faster to 100). Also the high-gear 20 to 40, 30 to 50, and 40 to 60 times were so much faster than the Cosworth that Motor Magazine did not publish them. The Avenger pulled hard from a little faster than 10 mph in high gear. The Cosworth Escort did not pull down low enough to even do the 20 to 40 high-gear test.

After the story came out, I asked why the “in gear” numbers had not been published. The answer I got was, “If the guys at Cosworth read this they will already be embarrassed enough that a two-valve pushrod engine was faster through the gears; do you want to embarrass them further?” I did not answer that, but I really wish those numbers had been published because this engine was not about outright peak numbers, but rather a huge power-band width and superb drivability.

So what produced the Avenger engine’s steam-engine-like low-speed torque, along with its race winning top end? It was not small intake valves as some testers suggested. It was not a reconfiguration of the engine’s big bore/short stroke to a small bore/long stroke. Nor was it countless other erroneous tales of moves that are supposed to produce a torquey engine.

The real-world factors that made this engine successful were as follows. First, the head flowed air very well, especially at low lift. And I used as much swirl as I could find, although it was still only about average in that respect. Second, its valves used almost the entire cylinder diameter. Because this was a short-stroke engine, the bore was big for a 1,500-cc unit, and that meant the valves were too. Third, and a very important factor on the list, the engine’s ports were the appropriate size. Great attention was given to the twin 2-barrel side-draft Dell’Orto 40 DHLA carb’s auxiliary venturis, to ensure a uniformly fine fuel atomization. Also the independent runner induction (one barrel per cylinder) and a well-spec’d exhaust system (in terms of lengths, diameters, etc.) provided exceptional induction and exhaust.

This total package produced, even at very low RPM, a highly combustible mixture in the engine’s quenchless chamber. Frosting on the cake came from a diligent calibration of the mechanical and vacuum advance delivered by the distributor. If I had to isolate one factor that contributed to this engine’s wide-ranging performance, it has to be its obviously superior in-cylinder mixture dynamics.

Sifting Through the Data

Let’s take stock of where we are. The first engine we looked at, the Mini’s A-Series, liked big lumps of fuel while the second, the Avenger engine, liked really small droplets. What this suggests is that there is not a clear-cut route to producing the best approach to optimum in-cylinder combustion dynamics.

At this point, you might wonder if the Chrysler Avenger’s quenchless-chamber engine was something of an enigma. Were we fixing some inherent shortcomings, such as the quenchless combustion chamber, to show results that were very positive on this engine but less likely to show on more conventional engines? Well, that could be. But if this engine was, so to speak, acting as a magnifying glass on combustion dynamics, it is still a good tool with which to work. However, later on, it was found that this seemingly odd-ball engine was not so far from a mainstream case as might have been first believed.

British Touring Car  Championship Year

After a dazzling show of speed (beating cars of twice the displacement) during the last few races of the previous year, Chrysler’s race boss, Des O’Dell, put his support behind the Vizard three-man team. He gave us a car and all the factory parts we needed to build a BTCC car, which is a  championship for a manufacturer’s title and is contested on an international level. It’s a bit like having NASCAR Cup Car racing with every major world manufacturer competing. We were up against twin-cam engines of Alfa Romeo, Lancia, Renault, and Toyota as well as the big bucks of Ford Motor Company, General Motors, and the like.

How did we do? We came out of the gate fast. By about the fourth race, our 2-buck, all-iron, push rod powered grocery getter was by far better than the competition’s cost no object, twin-cam sports specials. Did we win any races? Hell, no!

Our competition’s engines barely made 7,500 rpm. Our first engine of the year had a shift point of 8,800! By the middle of the season we were turning this pushrod engine to 10,500 and, between two corners at Brands Hatch, to 11,000 rpm. What that meant was during our test sessions (i.e., the race) we broke about one each of every part that could break. Sure, we had the fix by the following race, but that did not exactly help our cause that day.

Also we were running the races as part-timers. From Monday to Thursday, 8 am to 6 pm, we all had full-time jobs. From 7 pm until 1 or 2 am in the morning and Thursday through Sunday evenings, we either worked on the car or raced. Our budget for the year was less than what most teams spent per race. By the end of the year our team had managed a second place plus a couple of thirds, a class pole, and maybe half a dozen fastest laps. During six races the car had broken in a new engine in practice to put it on the grid on either the last row or the second to last row.

Sounds bad at this point, but we were learning and getting better all the time. The good part is that before the end of the first lap my Avenger was either first or second! I said the car was fast, and fast is exactly what I meant. So where did all this speed come from? Just as before, no one thing in general, but I can say that cylinder head flow was important. Low-lift flow (very important) was significant, along with cam design (critical), exhaust (big priority), and mixture characteristics and combustion dynamics (of great importance). Let’s start with mixture characteristics.

Weber Revamp

The homologated (that means the ones the as-manufactured car is supposed to have) carbs of the year for the Avenger being run were a pair of side-draft 40-mm Weber DCOEs. These came equipped with 30-mm main venturis. The rules allowed a change of main venturis for any design we wanted, but the hole could be no larger than 30 mm. Also the auxiliary (booster) venturi was free. This gave me scope to make new auxiliary venturis based on what I had learned from the Dell’Orto. The result was an average of about 5 ft-lbs of additional torque throughout the entire RPM range. Peak power, which occurred right around 8,000 rpm, was up by about 7 hp. This increased output was better than anything that could be built using off-the-shelf Weber parts. At this point, I concluded that I had achieved about as good a fuel atomization as the engine needed, so I turned to the cylinder head.


Fig. 11.10. There are always horsesfor-courses. As impressive as this Pierce Weber setup is for a small-block Chevy, the 40 DCOE Weber variant was far from matching the 40 DHLA Dell’Ortos on the quenchless chamber of a 1,600-cc Chrysler Avenger engine.

Fig. 11.10. There are always horsesfor-courses. As impressive as this Pierce Weber setup is for a small-block Chevy, the 40 DCOE Weber variant was far from matching the 40 DHLA Dell’Ortos on the quenchless chamber of a 1,600-cc Chrysler Avenger engine.

At the time, the BTCC rules and regulations called for a stock valve lift, which in this instance was a menial 0.390 inch. This meant intake valve acceleration and flow, especially from low lift, became vitally important. In fact, this is just one more incident where high flow at low lift won the day and did not, as is so often claimed, cost power.

F1 engine manufacturers Cosworth and Judd were my competition in the cylinder head department. I won’t go into too much detail here because combustion dynamics is the subject but, suffice it to say, the low-lift flow on my head was about 30 percent more at 0.050 than the Cosworth head while the full-lift flow was identical. Although there is more to it than just low-lift flow, it’s worth noting that my Avenger head, on this 1,600-cc engine, made 11 hp more than the Cosworth modified head!

Critical Port Finish

Probably because of its quenchless chamber and the need for a homogeneous and well-atomized fuel mix, the Avenger’s intake port size and finish was critical. In this instance I used a port that was a full 1/4 inch smaller in diameter than the ones used in the Cosworth or Judd heads. Also, unlike the fine finish of my competition’s heads, my ports were rough by virtue of a 40and 60-grit emery roll finish. This reduced the tendency of the fuel to coagulate and form rivulets prior to entering the cylinder. Notice I say it “reduced” the tendency; it did not cure it by any means—just made it a lot better. (See Chapter 4 for more info about wet flow and how to minimize it.)

Finally: The Chambers

After all the discussion on the carb’s air/fuel mixture preparation capabilities, let’s examine the port size and finish and the valve flow characteristics. The scene is set for us to look at the combustion chambers. The Avenger’s open chamber looked nauseatingly simple and I felt, since emissions were of no concern, it must be possible to do a better job in terms of power.

As it happens, the rules specified such things as valve size, compression ratio, etc., but did not specify combustion chamber shape. So I started by finding the heaviest pistons (there was a minimum-weight limit and factory-original pistons had to be used) and bringing them down to weight by machining the piston crown. This allowed the top ring to be nearer the piston crown, thereby cutting the ring land crevice volume. That little space is way more influential than you may suspect.


Fig. 11.11. There is more to the combustion chamber than that which resides in the cylinder head. Not only is it necessary to take care of business prior to mixture arrival at the cylinder but also after the charge is trapped within the cylinder.

Fig. 11.11. There is more to the combustion chamber than that which resides in the cylinder head. Not only is it necessary to take care of business prior to mixture arrival at the cylinder but also after the charge is trapped within the cylinder.

Fig. 11.12. The top ring land volume is also a crevice volume. Crevice volumes are bad news for power, mileage, and emissions and need to be minimized as much as possible.

Fig. 11.12. The top ring land volume is also a crevice volume. Crevice volumes are bad news for power, mileage, and emissions and need to be minimized as much as possible.

 

Along with the piston mod, I also investigated the chamber form for flow efficiency. On the flow bench, I found that better flow, especially at low lift, could be had by forming a shallow chamber around the intake and exhaust valves. This necessitated machining the top of the block to get back to the 9.9:1 (as I remember) CR called for. This move was done a step at a time from one build to another.

We were essentially building, for race and R&D combined, about one and-a-half engines per race. Each time a build or rebuild was done the chamber volume in the head was increased. And as the chamber increased in volume, the block volume was reduced by machining the block deck. Each time we did this the package more closely approached a conventional chamber with squish.

At each new spec, 0.020-inch more material had to come off the top of the block to bring the CR back up to 9.9:1, and each time we saw more power. When the situation got to where the piston was 0.080 down the bore and producing the best results to date, I decided that it looked worthwhile to go whole hog and put the entire combustion chamber into the head and deck the block for a tight quench, rather than possibly do four more builds to get there. The results on the dyno were just shy of startling. If all had followed previous form, I would have expected about 6 hp more from this combo; instead it was 8 fewer!

More Combustion Curiosities

If the love of big fuel droplets for apparently the best combustion dynamics with the Mini Cooper engines intrigued you, here’s another somewhat mystifying combustion-dynamics case concerning, again, a Mini engine. After a successful season with a 1,293-cc Mini Cooper S hill-climber (finished second in championship and only narrowly missed first spot due to a priority problem; driver/owner went off on a honeymoon and missed a round), my client asked if I would build him a blown, bored, stroked, and generally no-holds-barred version of this engine. I did just that—I stretched capacity to 1,442 cc and installed a large Shorrock supercharger. Just so that we did not have to use an intercooler (there was no room), I limited boost to about 12 psi.

Dyno testing was to be on a chassis dyno. The successful engine from the previous year made an even 100 hp at the front wheels. After a huge outlay of cash by the owner, a break-in period, and a change of oil and plugs to a race-grade type, the engine was given its first wrestling match with the dyno. This resulted in only 85 hp in spite of the 12 pounds of boost going into this big A-Series engine. This was embarrassing to say the least. The owner was standing right there watching and here my partner Mike Lane and I were, with a car that sounded for all the world like it could break the unlimited land speed record, yet it was way down on even a conservative estimate of what it should make.

Mike and I dove into a search for all the likely causes of this awesome-noise-but-no-power deal. We diligently checked ignition timing, valve lash, ignition box, etc. All were okay. At several points along the way we made dyno runs but with the same results—about 85 hp. Finally, we got all the way down to pulling the front end off the engine to check cam timing. It was right where we thought it should be.

During re-assembly, the engine was inadvertently flooded. With the engine virtually cold, it may not have fired up too well on those super-coldgrade plugs. If they were also wet it certainly wasn’t going to be that happy from a cold start. So we pulled the wet Champion race plugs and installed an equivalent heat range of dry Auto-lite race plugs.

When fired up, the engine sounded no more or no less wicked than before but the dyno numbers were—at 142 hp—nothing short of a techno shock. Although a pleasant surprise, it was very much a case of “What in the world is going on here?” This was such a surprise that the Champions were re-installed and re-tested. Same killer sound— just 85 hp!

I have to tell you that in just about every other Mini application the Champion plugs were as good as, or better than, anything we could find but here was an anomaly. This engine apparently did not like anything with Champion written on it. So with no further ado we went on, with the Auto-lite plugs installed, to finely tune the big carb on the engine and get the timing right on the money.

After three hours, we had the ignition and carb dialed in, and this engine ended up pumping out 170-plus-hp at the wheels. I had hoped for about 185, but that’s dyno testing for you—it’s a lens to focus on reality. This may not always be as gratifying as fantasy but, regardless of positive or negative results, you do get to learn a lot more about what it takes to win races. And even though it was less than what I had hoped for, this Mini won our driver the class championship.

Failure Highlighting?

So why am I highlighting these negative results? Simple. I want to emphasize that the subject we are dealing with here is anything but simple. I had no idea why power dropped in the instances I have just mentioned and here, with more than 45 years and tens of thousands of dyno tests later, and I still don’t have an answer.


Fig. 11.13. This spark form can be expected in a cylinder with low-mixture motion. In this instance, the required arc length in order to extend from one electrode to the other is about the width of the plug gap.

Fig. 11.13. This spark form can be expected in a cylinder with low-mixture motion. In this instance, the required arc length in order to extend from one electrode to the other is about the width of the plug gap.

Fig. 11.14. The spark form in a high-swirl combustion chamber. Note that the length of the arc is about 1/2 inch. If the ignition cannot fire a 1/2-inch gap, the cylinder fails to produce.

Fig. 11.14. The spark form in a high-swirl combustion chamber. Note that the length of the arc is about 1/2 inch. If the ignition cannot fire a 1/2-inch gap, the cylinder fails to produce.

In the Chrysler Avenger instance the results, in terms of quench clearance, were about 180 degrees opposed to all the other tests I have been involved with. This Avenger engine liked to have the piston stop 0.120 inch shy of the head face (0.080 down the hole plus a 0.040 head gasket) for best results. For just about every other engine I have tested like this, that piston-to-quench surface gap is about the worst in terms of low-detonation resistance and poor combustion. It really begs the question of whether or not we can give an engine too much or too aggressive a quench action. The answer is—yes!

Take a look at Figure 11.13, which shows a spark jumping a gap in a chamber with low mixture motion. You can see that the length of the spark is barely more than the plug gap width. If the mixture motion is high, the spark becomes distended because it cannot take a direct route from one electrode to the other. The result is the spark has to break down a longer path through the mixture to strike an arc. In other words, it’s just like the plug gap has become bigger. If the ignition system is able to fire the plug, the distended spark produced starts off a larger and more aggressive flame kernel. This is more effective because it lights off the charge faster and more effectively. If the spark is not of high enough voltage to strike an arc, a misfire results and power drops.

Atomization Optimization

It is easy to jump to the conclusion that the better the atomization, the better the power output. If only it were that easy! In reality it is far more accurate (but still not 100-percent true) to say: The better the fuel is finely atomized (and/or vaporized), the better the brake specific fuel consumption (BSFC) is. This number should not, as is so often the case (even with pro engine builders) be confused as an indicator of the mixture ratio. It is only roughly connected to the mixture and is in no way a measurement of mixture ratio. In other words, it’s only a consequence of the ratio.

At this point the question is: “Can the fuel be atomized and vaporized too much?” Let me set the scene. It’s about 1977 and I am just starting testing on some of the trick carbs built by Tucson’s premier carb builder/designer, Dave Braswell. The year before, at the 1976 SEMA show, I got to talk with Holley’s then–chief engineer Mike Urich. In our conversation I was amazed to find that, as far as Mike knew, Holley had done no official research on the effects of booster design on fuel atomization. I mentioned this somewhat surprising tidbit of info to Dave Braswell and he immediately volunteered assistance and carbs to do some testing on what we perceived as a typical street-tuned small-block Chevy. Here is how things unfolded.


Fig. 11.15. Here are some typical Holley boosters.

Fig. 11.15. Here are some typical Holley boosters.

Fig. 11.16. Since the Holley carb is used on more race engines than probably all other brands put together, I am going to use it as an example of fuel atomization versus booster (auxiliary venturi) style.

Fig. 11.16. Since the Holley carb is used on more race engines than probably all other brands put together, I am going to use it as an example of fuel atomization versus booster (auxiliary venturi) style.

Fig. 11.17. The degree of fuel atomization depends on numerous factors. The main tools to work with here are the booster’s gain, the air corrector size, and the hole pattern on the emulsion tube of the carb.

Fig. 11.17. The degree of fuel atomization depends on numerous factors. The main tools to work with here are the booster’s gain, the air corrector size, and the hole pattern on the emulsion tube of the carb.

The tests involved two carbs; each was about 750 to 800 cfm. One had high-gain fine-fuel-spray dogleg-style boosters, and the other had typical low-gain coarser-spraying straight-leg boosters.

The engine was run with three intake manifold types. The first was the stock exhaust, heated, and consequently hot-running intake. The second was an aftermarket two-plane aluminum intake with the heat crossover blocked off. The last was a Victor Jr. intake, which, as an airgap-style intake, was even more significantly cooler running.

On the stock intake, which was also the hottest by far, the trickedup Braswell carb with its small fuel droplets lagged behind the nearer stock carb by 8 hp (on a nominal 360-horse engine), but the fine-fuel spray delivered by the trick booster carb produced, by a small margin, the best BSFC both at wide-open throttle and part-throttle cruise. On the heat-blocked aftermarket two-plane intake, the carbs were very close in terms of output, but we are still considering a relatively hot-running manifold here.

The BSFC with the fine-fuel droplet booster carb was as much as 8 percent better, especially at part throttle. On the cool-running Victor Jr., the finely atomizing booster-equipped carb was unbeatable everywhere in the RPM range. It made about 12 hp more, and the brake specifics were all better (lower) numbers by a substantial margin.

So what does this tell us? The results indicate that there is an optimum fuel-droplet size that balances the need for some vaporization against the need to not evaporate too much fuel and spoil the engine’s volumetric efficiency (VE). Hot-running engines can offset the negatives of big fuel droplets from the boosters but cold ones cannot. Cold intakes need the ratio of the fuel droplet’s surface-area to volume ratio increased (which is just what happens as the fuel droplets get smaller), so that the loss of the heat as a vaporization source is compensated for by increases in the fuel’s evaporative surface area.

Atomization in Practice

In the early 1990s I became involved with booster development with the Carb Shop in California. The plan was to develop a Super Booster that not only gave a big signal but also did not obstruct the airflow of the main venturi to any greater extent than a regular high-performance booster. If a high-gain booster can be used it means that, for any given application, a bigger carb can be used for more top end before drivability and low-speed output suffer.


Fig. 11.18. The simple high fuel atomization mods I did to the Constant Vacuum Strom-bergs as fitted to the Chrysler Avenger GT. As simple as these mods were, they netted no less than 12 hp to this 1,600-cc engine output—mostly in terms of added torque.

Fig. 11.18. The simple high fuel atomization mods I did to the Constant Vacuum Strom-bergs as fitted to the Chrysler Avenger GT. As simple as these mods were, they netted no less than 12 hp to this 1,600-cc engine output—mostly in terms of added torque.

Well, the program produced some trick-looking high-gain boosters, which found their way (unknown to me) into the carb(s) of a front running NASCAR Cup Car team. On the dyno in the crisp December days just before Christmas in Mooresville, North Carolina (the ancestral home of all Cup Car teams), these boosters paid off in the team’s Daytona 500 engines to the tune of about 10 hp. So, with great expectations, the team headed off in early January for the Daytona 500 in Florida.

It was hotter than usual that January in a normally hot Florida. With the new carb, the car was well off the pace. In frustration, the team replaced the trick booster carb with the old one, and the car immediately ran on the money. The lesson here is that you can absolutely guarantee that too much of a seemingly good thing is—not so good. The percentage of fuel atomizing prior to entering the cylinder was such that any gains in better combustion were overridden by the drop in VE caused by the added fuel vaporization taking place within the intake manifold.

So when is a high-gain booster any good to a race car engine? Rarely, if ever, as things stood for a typical Cup Car engine of the early 1990s— but let’s move on.

Thermal Barriers

In the early 1990s, I was heavily involved in thermal barriers. It’s something I have looked at on and off since my Formula Ford days in the late 1960s. We found a temporary 2 hp (it barely lasted a race) by using high-temp exhaust paint on the pistons (Sperex, I believe). In this case, a relatively extensive study was made of the effect of thermal barriers in race intake manifolds.

Using a single 4-barrel carb on a single-plane intake, the effects of various boosters with the intake were explored both with the intake runners raw (uncoated) and then with them coated with a thermal barrier. The booster that worked best with the raw runners was of the stepped dog-leg variety shown in Figure 11.17. When the manifold was coated and used in conjunction with this booster, the power figures were within about a horsepower or so of unchanged.


Fig. 11.19. This thermal barrier coating added power for this race-winning head I ported for a 2-liter Mitsubishi engine, and it also extended its life from about 20 passes to more than 100.

Fig. 11.19. This thermal barrier coating added power for this race-winning head I ported for a 2-liter Mitsubishi engine, and it also extended its life from about 20 passes to more than 100.

So what’s the deal here? With the raw ports and mixture temps measured on the number-2 runner we saw, from a carb intake temp of 84 degrees F, a drop to 55 degrees F due to the evaporation of a portion of the fuel. By dividing the plenum front to back and using one end of the engine to drive the other (and no fuel to the front float bowl), we found that the air at number-2, without any fuel, picked up (allowing for a few corrections) about 10 degrees of manifold heat (more at low RPM and less at high). With the coated manifold, the engine’s number-2 intake runner temp was between 5 and 8 degrees less.

So the thermal barrier was doing what it was supposed to do, which was keep heat out of the intake charge. What was not happening here was any evidence of any increased power due to the cooler charge. When the charge temperature was measured on a functioning number-2 cylinder, the drop in temperature from the carb to the head/manifold interface of the  number-2 runner was only barely changed; and if everything had been working as before, it should have been at least 5 degrees cooler. This indicated that the cooler-running intake was not allowing as much fuel to vaporize and, therefore, the added wet fuel arriving at the cylinder was compromising the combustion process.

At this point, we installed high-gain annular discharge boosters along with appropriate (bigger) air correctors. With the same air/fuel ratio, the better atomization restored the percentage of vaporized fuel entering the cylinder. The temperature at the number-2 runner with an air/fuel mixture passing through dropped to 49 degrees F.

The dyno also showed some meaningful gains at this point. Essentially, the cooler-running, more finely atomized charge had the effect of moving the entire torque curve in an upward direction. On a nominal 450-hp engine, the torque at 3,000 rpm rose by 11 ft-lbs (6.3 hp) and by 7 ft-lbs at 6,200 rpm (8.3 hp).

Swirl and Quench

Swirl (and in the case of a four-valve engine, tumble) and quench both introduce mixture motion to the charge trapped in the cylinder. At the beginning of this chapter, I discussed the combustion speeds seen within a typical running engine and pointed out they were probably a lot slower than might be expected. This is where swirl, tumble, and quench begin to play a vital part. Just how vital is shown by a test I did many years ago that, looking back, I wish I had done on a more comprehensive level. But hindsight is always 20/20, so here is what I did learn. The test engine was the almost inevitable small-block Chevy. This was just a mule engine and had a relatively short cam, cast pistons, a 9.025-inch-deckheight block giving about a 0.060inch quench (a little on the wide side of optimal), and an 8.6:1 CR.

Flow testing the head on the flow bench, I found I could block off about 1/8 inch of the port on one side (cylinder-center side) and have good swirl, but doing the same on the other side (cylinder-wall side) produced poor swirl.


Fig. 11.20. The port test configuration for the high-swirl/low-swirl dyno test in Figure 11.21.

Fig. 11.20. The port test configuration for the high-swirl/low-swirl dyno test in Figure 11.21.

In each instance, the flow was about the same. I ran the tests with an exhaust-heated intake manifold with the heat on, to minimize any effect of any possible changes in wet-flow characteristics. On the 305 test engine, the torque output at 2,250 rpm (that was as low as I could go on the dyno) increased 30 ft-lbs with the higher-swirl port. As RPM rose, the two torque curves got closer. But in this instance, the low-swirl port configuration never got to be as much as the high-swirl. From this, you can see that at low RPM mixture motion is important.

It is also of interest that at low RPM (2,250) the best timing was as much as 5 degrees less with the high-swirl port/chamber than the low-swirl one. At peak RPM, the total timing was not that much different, with 30 degrees for the high-swirl and 32 degrees for the low-swirl.

As RPM increases, the situation can change. First, if an engine has a fuel system that is delivering somewhat larger fuel droplets than might be optimal, high swirl can centrifuge, or spin, the fuel onto the cylinder wall, thus depriving the main body of the charge of fuel. That places even more fuel in the crevice volume and a leaner-than-supposed mixture during the bulk of the combustion cycle. Neither scenario is good for output or fuel economy.

The lesson here is that higher swirl values in high-RPM engines need to be accompanied by good mixture preparation. When RPM numbers and piston speed get to those typically seen in race engines, the charge is substantially agitated, so the need to add mixture motion becomes less of an issue. However, my experience is that at least a small amount of swirl is a positive, even in a 10,000-rpm 500-ci Pro Stock engine. I can’t say if that carries over to a multi-valve F1-style engine. My first thought is probably not; but probably not many of you are designing and building them, so it’s not something I am going to worry about.

For a multi-valve engine, the natural tendency is for the charge to tumble as it enters the cylinder.

This is okay, but there is a difference in the way port-induced mixture motion reacts as the piston comes up the bore. By compressing a rotating charge into a combustion space of less diameter/area than the bore (as is usually the case), due to the law of conservation of angular momentum, the swirl can actually increase in value. In contrast, the higher the compression goes, the more the tumble motion is suppressed. At CRs of 11:1 or more, any tumble that existed when the piston was at the bottom of the stroke is largely gone. This is one of the reasons I worked so diligently to generate meaningful swirl on the Poly-Quad chamber head described in Chapter 10.

More Thermal Management

From the discussion above, it has become evident that thermal barrier coatings may have more to offer than might initially have been thought. But I have not finished this subject yet. I have worked with many of the leading companies, such as Swain Tech, Tech Line Coatings, and Poly-dyne Coatings over the years, but since about 2002, I spent a lot of time working with Calico Coatings. It has a great nearby facility and I can go to discuss whatever experiments I am into at that time. Without exception, the company has been ready to help in such tests and that has allowed me to move along on coating tests at a rate that was otherwise not possible. That is not to say those other companies have not contributed here. As this is being written, I am involved in some interesting experiments with Tech Line Coatings.


Fig. 11.22. If you have to pick one single element of the induction/exhaust of the head, coat the valves. Here, you see the Calico Coatings treatment to the intake and exhaust valves of one of my 5.0 Mustang racers. The biggest power influence is the face of the intake valve because it cuts heat going into the intake charge.

Fig. 11.22. If you have to pick one single element of the induction/exhaust of the head, coat the valves. Here, you see the Calico Coatings treatment to the intake and exhaust valves of one of my 5.0 Mustang racers. The biggest power influence is the face of the intake valve because it cuts heat going into the intake charge.

On the subject of in-cylinder combustion dynamics, we can say that for a given fuel, ambient weather conditions, and a host of other factors, there is a certain ratio of wet to-vaporized fuel that is optimum for best output. Based on everything we have discussed to date, we can say that keeping the fuel in suitably small droplets, allowing only a given percentage to vaporize, and avoiding wet-flow streams as far as possible is a major factor toward increased power from an engine (other than a Mini).

But before we wind up, there’s one more point to make. All the measurements of the intake charge temperatures were done at the intake-manifold to cylinder-head interface. But heated or not, the fact is the intake manifold is not usually the greatest source of heat input into the charge.

A case comes to mind here, which occurred about 1990 during some tests I did for Cosworth with a 2,000-cc BDP Midget sprint car engine. I had reason to turn off the dyno cell lights for a photo of the nearly white-hot exhausts seen during a run on this 280to 281-hp injected engine.

Being in the dyno cell with an engine turning 8,500-plus rpm can be a little unnerving, but as I walked past the deafening intake I realized that I could see, down the intake stacks, the intake valves glowing very dull red. At this point, it had not occurred to me just how hot the intake valves could get. But consider this: Because the intake valves are of greater area than the exhaust they pick up, during the combustion cycle at least, the intake produces more BTUs of heat than the exhaust. And where do they dump a whole load of this heat? You guessed it—they put it right into the intake charge.

This results in one positive and one negative. First the positive: The fuel is further vaporized before entering the cylinder, yet by how much I have little idea. The negative: The intake charge is heated and that is not so good. So I considered what the tradeoff was between these possibly competing factors. Sometime later we pulled the head on a duplicate engine and coated only the intake valves. A week later the results were in.

First, a look down the intake stacks in a darkened cell revealed that the valves were no longer visibly glowing. I fully realize that this is hardly a scientific way to measure the valve temperature, but observation was all I had. The fact they were no longer visibly glowing meant, at a good estimate, they were probably a hundred degrees or so cooler. Second, power figures were up from the low-280s to the mid-280s with the coated valves.

We can see an increase in power with high-pressure fuel injection and line-of-sight ports exiting into an open area of a combustion chamber and when there is no chamber wall directly in the path of the entering charge. Are the wet-flow dynamics of a typical two-valve V-8 head likely to deliver similar results? Or do we need to see greater fuel shredding at the seat and/or greater heat input to vaporize more of the charge, by this or other means, in order to get the benefits of cooler intake valves? That, along with crevice volumes, is what we examine next.

Small Crevice Volume— Big Consequences

In most instances, crevice volumes, such as the one between the top ring land and cylinder bore wall (see Figure 11.26 ), look insignificant, and that is why this power-robbing element of the combustion chamber is so easily overlooked. Crevice volumes are the stealthy thieves of those few horsepower that can so often make the difference between you winning the race or not. However, before we investigate the disproportionately negative effect crevice volumes can have, I need to establish exactly what defines a crevice volume.

The worst crevice volume is that contained within the piston’s top ring land. But this, is not the only crevice volume that we have to contend with. A combustion chamber with a sharp corner between the floor of the chamber and the chamber wall is not so good.


Fig. 11.23. The power difference between a high set ring (0.150 down from deck) and a low set ring (0.375 down from deck). All this indicates is that crevice volumes are far more influential than their small stature might suggest. To get results this concise, each combination was run seven times with the best and worst thrown out, and the remaining results were averaged.

Fig. 11.23. The power difference between a high set ring (0.150 down from deck) and a low set ring (0.375 down from deck). All this indicates is that crevice volumes are far more influential than their small stature might suggest. To get results this concise, each combination was run seven times with the best and worst thrown out, and the remaining results were averaged.

Fig. 11.24. For a detailed breakdown of what is happening here, refer to the text on this page.

Fig. 11.24. For a detailed breakdown of what is happening here, refer to the text on this page.

Some years ago, I watched a combustion cycle filmed through the wall of a quartz-cylinder engine. It was quite surprising to see that though there was about 1/8-inch radius between the chamber floor and wall, the charge that was immediately in that radius did not burn until much later in the cycle than did the bulk mixture adjacent to it. What this told me is that the floor-to-wall radius in any combustion chamber needs to be as large as possible.

There are also crevices between the edge of the valves and the combustion chamber or cylinder walls. These are the worst on a parallel-valve head, such as we see used in most small block Chevys, Fords, and Chryslers (but not the new Hemi). Although we need to keep these crevices in mind, the one we need to focus on the most is the top ring land volume.

At this point, you may be thinking that the ring land volume is so small compared to the volume of the rest of the cylinder that it cannot possibly have much influence on anything. At first, it may seem that way, but the opposite is actually closer to the truth.

For an example, consider the cylinder volume of a 350-ci small-block Chevy: 717 cc. To this, we must add the total combustion chamber volume. Assuming a 10:1 CR, this is 89.5 cc. The ring land volume of a typical off-the-shelf high-performance piston for this is right around 2 cc. At the bottom of the compression stroke, the crevice volume within the piston top ring land represents just 0.25 percent of the whole. That’s 1/4 of 1 percent!

Now let’s take the piston to the top of the bore. Here the volume is now 89.5 cc, and the 2 cc in the top ring land represents not just 1/4 of 1 percent but 2.4 percent. In simple terms, this means that almost 2.5 percent of the inhaled charge now resides in the top ring land crevice volume.

At this point, someone doubting the importance of the crevice volume contained in the top ring land might just start to concede that it’s important to minimize the ring land volume. But it is still less than 2.5 percent. You might say “this is insignificant, so let’s not fixate on it.” But combustion factors in this example are about to change for the worse, as I continue to explain.

As the piston comes up the bore, the ring land volume does not tend to fill with a combustible mixture of fuel and air, but rather a very fuel rich mixture. Any wet flow of fuel that was on the bore walls at the start of the pistons travel cycle up the bore is scraped off by the piston rings. The motion of the piston up the bore and the motion of air and fuel into the ring land volume, as compression takes place, tends to push everything into it rather than letting anything out. By the time the piston has reached the top of the bore, probably half of the ring land volume is raw fuel. As the piston approaches TDC, spark initiates the combustion phase.

Now take a look at Figure 11.24. This is a computer-enhanced version of the photos taken through the transparent walls of a research engine. This was hardly a high-performance engine, but what we learn from it is directly applicable. Using what we see here, we can look at what happens to the mass of burned mixture compared to the volume. First the spark hits at 45-degrees BTDC. Note how the “mass fraction burned” (red curve) progresses very slowly at first. In fact 40 degrees after the spark has fired only about 5 percent of the charge mass has burned.

But take a look at the flame volume in the chamber at TDC. You can see that a lot more than 5 percent of volume has burned. What is happening here is that as the mixture burns, it expands and pushes the remaining unburned volume into a smaller space. This means we have the same type of scenario as we had when the piston was coming up the bore on the compression stroke. As the cylinder pressure rose, a greater amount of charge mass ended up in the top ring land volume. At TDC only 10 percent of the mass has burned, but the burned mixture volume takes up between 60 and 70 percent. Assuming it was all a homogeneous mixture in the cylinder, at TDC, we then have something on the order of 6 percent of the charge now in the ring land volume. That is 2,400 percent more than what was in it when the piston was at BDC!

The mass burned versus the volume burned does not catch up until about 20 degrees after TDC. At this point, the mixture that was compressed into the ring land volume starts to come back out and burn. Also the fuel that was trapped there is released as the piston accelerates down the bore. The mixture that burns is doing so too late in the cycle to contribute a proportionate share of the total energy released. If the entire 6 percent of the mixture that could be contained in the ring land volume were to be burned, it would only contribute about 2 percent of the resulting total energy.

Gas-Ported Pistons

So far we have looked only at regular-style pistons as opposed to race pistons with gas porting to the back of the top ring. Gas porting of pistons does not seem to have caught on in Europe, so for the benefit of those readers, check out Figure 11.25 showing the two common types of gas porting.


Fig. 11.25. The two common types of gas porting to aid top ring seal. The radial gas ports are the most popular, especially for endurance race engines; they have a significantly less tendency to clog and show less ring/bore wear than the vertical gas ports.

Fig. 11.25. The two common types of gas porting to aid top ring seal. The radial gas ports are the most popular, especially for endurance race engines; they have a significantly less tendency to clog and show less ring/bore wear than the vertical gas ports.

The idea behind gas porting is to seal the ring against the bore more effectively. Gas porting does work, but it also increases the crevice volume. So we have to make a thoughtful tradeoff when designing pistons. For what it’s worth, I favor radial gas ports because they never plug up with carbon deposits.

Piston Dishes

At this juncture, I should point out that as much as, or more than, half the combustion-chamber area resides in the piston. If the piston is to have anything of a cutout other than for the valves, it seems most logical to have the dish located as near as possible, right under the spark plug. As obvious as that may seem, things don’t always work out that way and often it’s down to dyno testing to see what works.

A test showed that the A-Series Mini Cooper engine preferred a valve cutout combined with the piston chamber that was no more than a plain slot—about 0.080 inch deep right across the piston. This worked better than conventional valve cutouts and a bowl under the plug, to the tune of 4 hp within 135 hp. Working with Sealed Power many years ago, I had the idea that a dished Chevy piston should have a sloping dish with the bulk of the dish right under the spark plug. This would, in effect, put more of the charge mass in closer proximity to the plug. It seemed such an obvious move but the 400-horse mule lost more than 10 hp with this style of piston. This shows again that things are not always as they seem when it comes to simple answers for chamber dynamics.

Testing Theory

So far we have mostly looked at the potential power loss from the prime crevice volume. The dyno is now the reference point. First let’s consider the eradication of the top ring land volume. During the mid 1970s, the primary topic in the auto-motive world was the problem emissions were causing in terms of power and fuel economy. In the late 1960s, Sealed Power came out with the Head Land ring. This was a top compression ring that eliminated the top ring land volume. Sales of this ring were really lackluster until the promises of lower emissions prompted a semi-revival in interest. See Figure 11.26.

About 1977, I had the opportunity of doing a back-to-back test between conventional Sealed Power top rings and Sealed Power Head Land rings. My co-conspirators on this test managed to borrow a three-gas analyzer, so we could look at some of the emissions with each style of ring. I have long mislaid the test results, but they were significant enough to stick in my mind. We saw a drop of at least 30 percent of unburned hydrocarbons and a drop of at least 20 percent in carbon monoxide. Since the rings primarily targeted emissions, that was good.


Fig. 11.26. The Sealed Power Head Land ring. Note that it gives a crevice-free piston top. The problems here for a high-output engine are ring mass and the encroachment of valve cutouts on the ring.

Fig. 11.26. The Sealed Power Head Land ring. Note that it gives a crevice-free piston top. The problems here for a high-output engine are ring mass and the encroachment of valve cutouts on the ring.

For power production, things were less clear cut. The bottom line here was that low-speed output went up. At about 1,800 rpm (that was as low as the dyno pulled down the test mule), torque was up by a solid 3 percent. At about 3,500 rpm, both styles of top ring were even, and at 5,000 rpm the Head Land ring torque was down about 2½ percent. As it happens, we had a simple blow-by gauge hooked up to the crankcase, and this showed that above about 4,000 rpm the Head Land ring was losing its ability to effectively seal. I talked to my friend and then Sealed Power’s chief engineer, Cal DeBruin, and he confirmed what we suspected. The mass of the ring and its large face was difficult to control at higher RPM and resulted in an increase in blow-by.

As conclusive as that may seem, it was not the end of the story. Enter Dale Fox, who worked with Smokey Yu-nick on and off from around 1968 until a few years before he passed in 2000. During 1969 and 1970 Dale worked full time for Smokey, and did a lot of whatever engine building Smokey could not personally find the time for.

Here is how it went for the second round of our look at Head Land rings. For the tests, the forged pistons used Sealed Power’s recommended clearances, which turned out to be about 0.005 to 0.006 inch. The width of a Head Land ring meant that to accommodate any rocking of the piston in the bore, the face against the cylinder wall had to be a barrel shape. Dale and Smokey liked the concept of the Head Land ring and, although their first round of testing showed pretty similar results to those I ran, they stuck with the concept. What they did was to see if they could cut piston clearances to a minimum, and thereby reduce the rocking of the piston in an effort to help seal everything. Their efforts paid off; they managed to get the ring to work much further up the RPM range. Remember, back then most endurance engines did not turn much in excess of 7,500 rpm.

So Far, What Have We Proved?

At this point we can say that eliminating the top ring groove crevice has proved that output can benefit. Other than minimizing clearances as Dale Fox and Smokey Yunick did, it is not quite obvious how we can implement the Head Land ring principle in an extreme high-RPM engine. Maybe, if the ring was made of titanium, weighing no more than a conventional steel ring, it would work up to a much higher RPM. Yes, I realize that someone is going to say that titanium is the most seize-able material on the planet and barely stops short of welding itself to ice when it’s rubbed against it.

The good news, if anyone is interested, is that I know how to make titanium into a better bearing material than phosphor bronze. This has allowed the production of Ti wrist pins (gudgeon pins if you live in England) that have no coating, are about twice as seize-resistant as tool steel pins, and have a Rockwell hardness of about 76. But I digress; so far I have not been in a position to do anything with respect to a Ti ring, but there are some moves employing more conventional approaches toward the reduction of the top ring land volume.


Fig. 11.27. The trend over the years has been for piston rings to become thinner and have less radial depth. Also, duty cycle permitting, the ring pack as a whole has moved toward the piston crown.

Fig. 11.27. The trend over the years has been for piston rings to become thinner and have less radial depth. Also, duty cycle permitting, the ring pack as a whole has moved toward the piston crown.

Sometime in 1978, I was at lunch with Pro Stock and engine building legend Bill “Grumpy” Jenkins, and the conversation turned to the subject of the top ring’s position in relation to the crown of the piston. Grumpy told me of the differences he had seen on a Pro-Stock-style, small-block Chevy. After my experience with the Head Land ring, I put a back-to-back test of a low-set top ring versus a high-set at the top of a long to-do list.

The problem was I needed to get a piston manufacturer to make two sets of pistons, for free and for no other reason than to establish what the difference might be. This opportunity came about when Moe Mills decided to leave Arias Pistons and, with a partner, start Ross Racing Pistons. Here, Moe made me two sets of pistons with valve cutouts minimized for the 280-degree (offthe-seat-duration) cam that was to be used.

The valve cutout depth is an important factor for an inclined valve engine such as a small-block Chevy because the cutout can easily intersect with the ring groove. To get that top ring up as far as possible, it was necessary to make sure the intake valve pocket was no deeper than needed. The low-set top ring was 0.375 down from the piston crown while the high-set one was 0.150. This gave crevice volumes of 1.36 and 0.55cc, respectively, for a reduction of 55 percent in crevice volume. Remember, we are only looking at a reduction of 0.81 cc, and you may well ask just how much a small change like that can possibly make.

The CR for the test engine was a solid 10.5:1. In this test, the pistons had a skirt profile that was intended to minimize piston-to-wall clearance. At the open end of the pistons, the clearance was only 0.001, which is relatively close for a forged piston. The clearance at the pin was 0.0035.

Also I need to make it clear that when I run tests I go to a lot of trouble to calibrate the carb right on, so there is a minimum of excess fuel. That is no more than whatever the engine wants for power optimization of the air/fuel ratio.

Additionally the carb had high-gain boosters (Braswell carb) that delivered good atomization and the heads were a relatively small port (165 cc) variety (ported factory 186 castings). All this should add up to a better-than-average quality of mixture arriving at the cylinders (i.e., minimum of wet flow versus airborne fuel flow). The tests were run after these pertinent parameters were taken care of. The numbers in Figure 11.23 are from an average of seven runs, with the best and worst tests thrown out in each case and the rest averaged.

This is a test done as diligently as possible with the equipment available. The situation for the “before” test was as good as it gets, with the engine showing very good brake specific fuel consumption. This indicates that the combustion process was already good. That probably meant the minimum of wet fuel entering the cylinders and, consequently, the minimum raw fuel into the ring land volume. Even with this best-case scenario, the power trend was decidedly upward.

Talking to various piston manufacturers has indicated that similar tests done on all-out race engines with a much higher CR have often shown bigger percentage gains. This makes sense; as more charge is driven into the ring land volume the higher the cylinder pressures go. From this, we can reasonably conclude that what is shown in Figure 11.23 is about the minimum increase that can be expected. In round numbers, we are looking at an average of 1.5 percent for some additional cleanup on an already good combustion process.

This is a pretty good result, because all the theory on how a nearly unprecedented amount of charge can end up in the ring land volume has not as yet accounted for one important aspect: The reason the top of the piston is smaller is that this is where the most heat is and the top thus expands more. What may be a 2-ccvolume cold may only be a 1-ccvolume hot. So, with the volume even smaller than we calculate it to be, we can see that a crevice volume that is even very small has a disproportionate effect on the combustion process.

Where from Here?

Other than setting the ring up as high as possible or getting a ring company to re-investigate the Head Land ring, what else can we do to rid the combustion space of the top ring land volume? About 20 years ago, Cosworth made a move to do just that. The company made the top land larger than normal, and then grooved it. The idea was that, during break-in and eventual WOT operation, the piston would touch the bore and wear the peaks from the grooving, so that the piston almost exactly fit the bore, thus eliminating most of the top ring land volume.


Fig. 11.28. This JE piston features many of the design parameters we have discussed as being pertinent to the combustion dynamics exhibited in a high performance engine.

Fig. 11.28. This JE piston features many of the design parameters we have discussed as being pertinent to the combustion dynamics exhibited in a high performance engine.

I remember talking to Cosworth’s piston designer (the late Geoff Roper) about this and the fact they were just starting to do this for Cup Car pistons. That move might just have been the start of Cosworth’s downslide in Cup Car piston sales! I forget what Cosworth called the grooving, but these days we commonly refer to them as “anti-detonation” grooves. How they help suppress detonation has not been convincingly explained to me, so I am still waiting on that one! I have never done a back-toback test for power increase either, but they have now been around for more than 20 years. I think the jury is still out on the advantages of anti-detonation grooves.

One aspect that creeps in here is the way increasing the cylinder pressure, either from compression or the combustion process, drives the charge into the top ring land volume. If you think about this, igniting the charge from the middle of the cylinder is the worst way to do this. If the charge burn is propagated from all around the circumference of the cylinder, the highest pressures occur long after the charge in close proximity to the ring land volume burns. Granted, this sounds like a good way to go, but firing the charge from all around the outside looks difficult to accomplish. It’s difficult, yes, but not impossible.

About 1995, I met a guy in California who had designed plug electrodes into a head gasket. The gasket had about 6 points around the circumference that fired the charge. Fel-Pro apparently liked the idea—at least initially—because it acquired the rights to it. I never saw any test data, but I heard it worked pretty well. Unfortunately production was not practical; the heads had to be removed to change the plugs!

 

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Conclusions

In this chapter, several things have become more clear. First, we talked about a lot of factors that influence combustion dynamics before even arriving at the combustion chamber. From this, it is evident that what happens inside the cylinder can be greatly influenced by how the intake charge is prepared prior to its arrival at the cylinder. Also the thermal barrier cases put forward here are examples from maybe a dozen or so tests. All were similar in their intent, and all show that the so-called Dyno Test Rule Number-1, concerning making only one change at a time and so often touted by doit-yourself performance magazines, is seriously flawed.

So, What Are the Rules?

From the foregoing, we can draw a reliably tentative conclusion about combustion dynamics. Namely that applying any universal rules can all too often turn around and bite your rear end. Just to keep my feet on the ground, I have a policy: All lessons learned, no matter how certain or obvious they may seem, should be regularly re-appraised for anomalies that may actually reveal something that was overlooked or misinterpreted.

What can be said is that paying attention to port velocity to give good swirl pays off, especially at low speed. But above about 8,000 rpm or so, swirl becomes academic or indeed may, under some circumstances, prove to be a disadvantage. Keeping the combustion chambers compact and as crevice-free as possible is a decided plus. Placing the plug in a high-mixture-motion area is also a plus. To this, we can also add that under most circumstances holding the quench tight is also good, especially for engines that are required to deliver over a wide range with, say, about 8,000 rpm as a max.

Raising the CR increases torque, and consequently power, throughout the RPM range. Because raising the CR increases thermal efficiency, it brings about an increase in fuel economy. If a longer-duration cam is installed, raising the CR at the same time can be worth considerably more than these two moves done separately.

When the CR is raised, peak combustion pressures are increased. A rule for typical production engines is that combustion pressures are equal to the CR times 100. This tells us that, from a 10:1 engine, we expect to see about 1,000 psi of peak combustion pressure. For a well-developed, high performance engine, combustion pressures can be as much as the CR times 120.

Written by David Vizard and Posted with Permission of CarTechBooks

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