David Vizard's How to Build Horsepower (S-A Design) [Illustrated] 1934709174, 9781934709177

Table of contents :
Title Page
Copyright
Contents
Introduction
Chapter 1: The Basics of Every Known Speed Secret
The Dynamometer
When Isn’t Bigger Better?
Chapter 2: Primary Point Induction
Air Temperature
Ram Air: the Basics
Practical Density Issues
Significant Pressures and Flow
Element Clogging
Chapter 3: Fuel Delivery Systems
Mixture Ratio
Exhaust Pollutants
Mixture Quality
Carb Function Basics
WOT Calibrations
The Boosters
Idle and Transition Circuits
Acceleration Enrichment
Carbs—How Big?
Carbs Versus Fuel Injection
Chapter 4: Intake Manifolds
Length and Volume
V-8 Intake Manifolds
Tunnel Ram Intakes
Chapter 5: Supercharging
Supercharger Types
Turbos and Centrifugal Turbine Superchargers
Boost Curve Shapes
Supercharger Selection
Superchargers and Built Motors
Getting the Cam Right
Turbo Cams
Fueling System
Chapter 6: Cylinder Heads
Optimizing Cylinder Head Airflow
Valve Shrouding
Practical De-shrouding
Ports
Port Evolution
Cross-Sectional Area
Applied Basic Porting
Wet-Flow Testing
Compression Ratio
Chapter 7: Porting and Flow Testing
Building a Really Trick Low-Cost Bench
Flowing the Exhaust
Porting
Results
Chapter 8: Ignition Systems
Number-One Ignition Goal
Spark Plugs
Coils
Ignition Timing and Curves
Modified Motor Timing Requirements
Modified Ignition Systems
Expected Results
Chapter 9: Real Camshaft Science
Mechanical Attributes Simplified
Hi-Perf Four-Cycle Engine
Airflow Dynamics
Choosing a Cam
Chapter 10: Cam Event Criteria
Overlap
Optimal LCA
Duration
Valve Lift: How Much?
Chapter 11: Valvetrain: The Physical Build
Reading Cam Spec Data
Valvesprings
Tappets
Timing Gears
Cam Timing
Rockers
Chapter 12: The Short-Block Engine
The Block
The Rotating Assembly
Connecting Rods
Pistons
Rings
Lubrication System
Chapter 13: Exhaust Manifolds
Pipe Diameter
Alternative Configurations
Using This Header Tech
Chapter 14: Mufflers to Tail Pipes
Simple Steps to Success
Cats and Mufflers
Muffler Flow: How Much is Needed?
Pressure Waves
Crossover and Balance Pipes
The Ultimate System?

Citation preview

CarTech®, Inc. 39966 Grand Avenue North Branch, MN 55056 Phone: 651-277-1200 or 800-551-4754 Fax: 651-277-1203 www.cartechbooks.com © 2010 by David Vizard All rights reserved. No part of this publication may be reproduced or utilized in any form or by any means, electronic or mechanical, including photocopying, recording, or by any information storage and retrieval system, without prior permission from the Publisher. All text, photographs, and artwork are the property of the Author unless otherwise noted or credited. The information in this work is true and complete to the best of our knowledge. However, all information is presented without any guarantee on the part of the Author or Publisher, who also disclaim any liability incurred in connection with the use of the information and any implied warranties of merchantability or fitness for a particular purpose. Readers are responsible for taking suitable and appropriate safety measures when performing any of the operations or activities described in this work. All trademarks, trade names, model names and numbers, and other product designations referred to herein are the property of their respective owners and are used solely for identification purposes. This work is a publication of CarTech, Inc., and has not been licensed, approved, sponsored, or endorsed by any other person or entity. The publisher is not associated with any product, service, or vendor mentioned in this book, and does not endorse the products or services of any vendor mentioned in this book. Edit by Paul Johnson Layout by Chris Fayers ISBN 978-1-61325-256-7 Item No. SA357 Library of Congress Cataloging-in-Publication Data Vizard, David. David Vizard’s how to build horsepower / by David Vizard. p. cm. Includes bibliographical references. 1. Automobiles—Motors—Modification. 2. Automobiles—Performance. I. Title. II. Title: How to build horsepower.

TL210.V5184 2010 629.25—dc22 2009044104 Written, edited, and designed in the U.S.A. 15 14 13 12 11 10

On the Title Page: The dyno validates or disproves theory. Time spent here is worthwhile, but only if what is going on is largely understood. If results are misinterpreted, then much of what might be gained is forfeited. On the Back Cover Top Left: One of the Dominator carbs has been removed to show the internals of this Ultra Pro Machining ProStock-style intake. When fitted to a two-valve pushrod engine, power eclipsed the 165-hp-per-liter mark. All-out ProStock engines, as of 2010, are right around 175-hp per liter. Top Right: Here are the springs I favor for hydraulic applications. Note the small-steel beehive retainer is as light as a titanium one for a regular spring. Middle left: Shown here is the real-life lobe centerline angle for a small-block Chevrolet roller follower cam. Middle right: The “shear dam” to the right of the intake guide boss can be clearly seen here. A fuel rivulet arriving at the shear dam is sheared off and re-introduced into the air. From here, it mostly excites the valve in the 6 to 7 o’clock position, and in doing so, aids swirl production. Bottom left: The block is the primary component for any engine. The block needs to be strong and machined to fine tolerances if winning results are to be achieved. Bottom right: The intake ports of the cylinder head shown here are the subject of an intense flow-bench development program. The high-efficiency figures produced made it possible to develop more than 100-hp per liter from a 355-ci (5.8L) 10.5:1 CR engine running service-station 92-octane premium fuel.

OVERSEAS DISTRIBUTION BY: PGUK 63 Hatton Garden London EC1N 8LE, England Phone: 020 7061 1980 • Fax: 020 7242 3725 Renniks Publications Ltd. 3/37-39 Green Street Banksmeadow, NSW 2109, Australia Phone: 2 9695 7055 • Fax: 2 9695 7355

CONTENTS

Introduction Chapter 1: The Basics of Every Known Speed Secret The Dynamometer When Isn’t Bigger Better? Chapter 2: Primary Point Induction Air Temperature Ram Air: the Basics Practical Density Issues Significant Pressures and Flow Element Clogging Chapter 3: Fuel Delivery Systems Mixture Ratio Exhaust Pollutants Mixture Quality Carb Function Basics WOT Calibrations The Boosters Idle and Transition Circuits Acceleration Enrichment Carbs—How Big? Carbs Versus Fuel Injection Chapter 4: Intake Manifolds Length and Volume V-8 Intake Manifolds Tunnel Ram Intakes Chapter 5: Supercharging

Supercharger Types Turbos and Centrifugal Turbine Superchargers Boost Curve Shapes Supercharger Selection Superchargers and Built Motors Getting the Cam Right Turbo Cams Fueling System Chapter 6: Cylinder Heads Optimizing Cylinder Head Airflow Valve Shrouding Practical De-shrouding Ports Port Evolution Cross-Sectional Area Applied Basic Porting Wet-Flow Testing Compression Ratio Chapter 7: Porting and Flow Testing Building a Really Trick Low-Cost Bench Flowing the Exhaust Porting Results Chapter 8: Ignition Systems Number-One Ignition Goal Spark Plugs Coils Ignition Timing and Curves Modified Motor Timing Requirements Modified Ignition Systems Expected Results Chapter 9: Real Camshaft Science Mechanical Attributes Simplified Hi-Perf Four-Cycle Engine

Airflow Dynamics Choosing a Cam Chapter 10: Cam Event Criteria Overlap Optimal LCA Duration Valve Lift: How Much? Chapter 11: Valvetrain: The Physical Build Reading Cam Spec Data Valvesprings Tappets Timing Gears Cam Timing Rockers Chapter 12: The Short-Block Engine The Block The Rotating Assembly Connecting Rods Pistons Rings Lubrication System Chapter 13: Exhaust Manifolds Pipe Diameter Alternative Configurations Using This Header Tech Chapter 14: Mufflers to Tail Pipes Simple Steps to Success Cats and Mufflers Muffler Flow: How Much is Needed? Pressure Waves Crossover and Balance Pipes The Ultimate System?

INTRODUCTION As I type this I am approaching my 52nd year modifying engines—the last 45 doing so in a professional capacity. I would like to say they have all been smooth but that is far from the reality. It was in England, my home country, and in 1966 that I had my first taste of building/modifying a V-8 engine. It was a Ford Flathead, the principal components of which I lucked into as an unfinished project. The gentleman I bought these cut-rate parts from was to be the guest of Her Majesty’s Government and so had no need of them for 10 years or so. A thick-wall block with a big overbore and a billet stroker crank resulted in a massive displacement, for a Ford Flathead of 301 cubes. A lot of porting work and the use of high compression, a race cam, triple Strombergs, and a whole lot of other high-performance moves resulted in 224 hp and 300 ft-lbs as measured on a dyno of unknown accuracy. Up to this point I had used only chassis dynos to prove the validity of my work. This was my first go around on an engine dyno and, as it has subsequently proved to be, the first of more than 250,000 dyno pulls over the next 42 years. At the time I was disappointed with the results from that Flathead, but 30 years later I was to find that this output was a relatively creditable result for a side-valve engine. But the disappointment of that engine was to be offset a couple of years later. Over the winter of 1968–1969, I was asked to port some small-block heads for a 302 small-block Chevy powering a Lola T70 sports racer. It seemed the heads used on the popular Bartz or Traco USA-built engines of the day had a propensity for cracking, and a new set of heads at a couple thousand dollars each (remember this is 1969 and a grand is more like $2,500 as of 2010) were needed for every race. I started with castings that I thought might do the job but were different from those used by engine builders in the United States.

Here I am about to time-in the cam on a low-buck 5.0 Ford Mustang engine. What is being passed on to you in these pages is 50-plus years of experience building high-performance engines. After a few weeks of experimentation, on my very simple flow bench, I had some promising working heads. The dyno subsequently demonstrated this very much to be so. That, in conjunction with a few other moves, proved to be the formula for killer results. With what I learned helping out on that engine, I reasoned (with good cause) that I had a working handle on what it takes to build an international-level race-winning small-block Chevy. Over the next few years I was to amass hundreds of wins, lap records, and pole positions, plus a half-dozen or so championship wins— two at international level—all with European engines.

What you see here is my 145-mph supercharged GMC truck. This is what I use to tow my racers, two of which you see here. The point I am trying to press home is that I actually live what I write. Fast forward to 1976–1977 in Tucson, Arizona, where I was industriously working on a Chevy rebuild book. As a side project, I built a hopped-up street 350 and looked forward to a similar degree of success as I had with the 302 in 1969. I had read and absorbed a huge amount of info from the big-circulation performance magazines and the product tests they published. I knew how to build a high-dollar race engine so I assumed that building a relatively high-output street engine on a more mundane budget would present little in the way of a challenge. Back then I was naive enough to assume that everything in black and white was a printed version of reality. With the parts I bolted together, I expected to have an engine of at least 375 hp and around 400 ft-lbs of torque. What I got was a pitiful 281 hp and something less than 340 ft-lbs of torque; and it had cost what was, for my budget, a sizable amount of cash for a “non-event” engine. A few phone calls to some engine builders who made their living building and dyno testing successful small-block Chevys caused me to re-evaluate much of the component choices I had made. It seemed my piston selection was far from optimal as was, among other things, the cam, carb, and intake manifold. With the input of sound advice, a return to the drawing board, and some more work on the flow bench, a new 9.5:1 pump-gas-burning street 350 was born. It was with a great deal of trepidation that I hit the dyno with this one. However, I need not have worried; the dyno delivered a respectable 402 hp

and 404 ft-lbs of torque. All this came with a smooth 650-rpm idle and more than enough vacuum to run power brakes.

Making 402 hp on pump gas was good in 1976 but things have moved on. Here is a 10.5:1 pump-gas 355-ci (5.8-liter) small-block Chevy street driver. Built on a reasonable budget with 2008 know-how and parts, it made 584 hp (more than 100 hp/liter) and turned 8,000+ rpm.

This 340-hp 5.0 Mustang engine was built on a very tight budget. The only new parts used were the intake manifold, MSD distributor, and the exhaust manifold (or headers). At the end of the day, the final analysis shows conclusively that knowledge is a key ingredient to building successful performance engines. That knowledge is best acquired from people who have hands-on experience developing successful engines. Acquiring such knowledge means the cash element in a successful build is vastly reduced. Here I have to say that, given the knowledge, you can cut the cost of a highperformance engine in half with ease, and sometimes cut it as much as 65 to 70 percent.

The dyno validates or disproves theory. Time spent here is worthwhile, but only if what is going on is largely understood. If results are misinterpreted, then much of what might be gained is forfeited. Unlike many performance books, this is not written by a journalist reporting on what others might think or have you believe they do to produce power. Think about this: Why would any successful builder/development engineer want to give away any real info that could well have cost them time and money to accrue? This book is a firsthand rendering of experience from 50-plus years of hands-on building for speed.

Information Transfer Logic Life has shown me that it’s one thing to have the knowledge to do something well, but an entirely different thing to pass on that knowledge in a simple and clearly understandable manner. Teaching is a science in itself

and, since the subject we are dealing with has many interacting complexities, it becomes very important to convey information in as logical a fashion as possible. My first thoughts here were to start at the point of the biggest impediment to power a high-performance engine has—the intake valves. But once I had covered that, which way should I go—up through the intake tract to the air filter or down through the exhaust? Adopting this approach, it seems that any organized logic failed at the second step so I rethought what might work best. The first move, within the pages of this introduction, is to make sure that you fully understand that the intake valves’ flow capability is by far the greatest impediment to the production of horsepower. With that understood, I tackle the production of power by starting at what seems to be the most logical point—where air enters the engine—and carry right through the system to the point where it exits. Somewhere about the middle of the book I deal with the short-block and considerations within it that affect power. Hopefully, by presenting the information this way, you can quickly look up anything relevant in a speedy and effective manner.

A straight pipe, or venturi, such as used in a carb, is an efficient means of transferring air from one place to another. On the other hand, a valve

(even when fully open) is not good because the route the air has to take to navigate around it is tortuous. In addition to this, it spends most of the time closed. When it is called upon to open, it takes time to reach sufficient lift to be able to deliver relatively high flow rates. The percentages here show typical flow efficiencies for each section of a typical production V-8 cylinder head intake port.

CHAPTER 1

THE BASICS OF EVERY KNOWN SPEED SECRET A bold chapter title to be sure but, before going into the fundamental principles of what makes horsepower, let us look at the definition of horsepower. This seems a necessary step because, too often, novice enthusiasts are unaware of how horsepower is defined and derived. The definition of 1 hp was determined by British engineer James Watt of steam engine fame. After suitable tests, he set the value of 1 hp to be 33,000 ft-lbs per minute. That means either lifting 1 pound through 33,000 feet vertically in one minute, or 33,000 pounds through 1 foot in one minute, or any combination that multiplies out to 33,000. From this, we can say:

where “work done” is the force involved (in pounds) and the distance (the number of feet) that it acts over in one minute. This is expressed as foot-pounds of work done per minute (not to be confused with ft-lbs of torque).

For a professional engine builder, a dynamometer is an essential tool to have. For the serious enthusiast, it is an essential tool to rent. But an engine does work rotationally so we need to translate the work done in a circle to the equivalent of a straight line. To do that, we multiply the crank radius (R) in feet by PI (3.142), then 2, and then multiply that by the force (F) in pounds as generated by the mean cylinder pressure and the revolutions per minute (RPM) involved.

But R × F equals Torque, so we can simplify this equation to:

In other words, we can divide the 2 × 3.142 × Torque × RPM on the top line by the 33,000 on the bottom line. This gives us (assuming English units) the universal formula for horsepower, which equals:

The Dynamometer

So now that we know how horsepower is derived, let’s look at how we actually measure it. For this we use a device known as a dynamometer. A dyno, which applies a braking force to the engine (hence, the term brake horsepower), does not directly measure horsepower but derives it by measuring the torque the engine can make at a particular test RPM and then calculating the power from these two known values. Power (as opposed to horsepower) is simply torque × RPM. To express the result in units of horsepower, it has to be divided by a constant. From the previous formulas, you can see the constant involved is 5,252. For example, a power calculation looks like this, if an engine produces 200 ft-lbs at 3,500 rpm. The formula is:

Here is how an engine develops torque. As the piston goes down, the bore pressure drops and the torque follows suit. Multiple cylinders, the mass of

the engine’s internals, and a flywheel tend to damp-out torque fluctuations. Double the torque at that RPM, and the power doubles. Doubling the RPM, but without changing the torque, also doubles the power. Therefore, you can see that any change in torque at a given RPM also changes the horsepower at that RPM.

1-1. Here, you see the basics of every speed secret known to humankind, and plenty that are not. It covers all the basic elements we need to improve to build more power. In the following chapters, we tackle each of the elements shown in more detail, starting with the factors at the top of this chart and finishing at the bottom.

Engine torque is exactly the same as the torque produced with a torque wrench (Force × Radius). In the case of an engine, the radius is half the stroke and the force is the average produced by the gas pressure pushing down on the piston. The gas pressure is the result of filling the cylinder with air and then heating it by burning fuel in it. Here it is worth taking a mental note of the fact that we only have atmospheric pressure (14.7 psi at sea level) to work with to fill a cylinder. This factor puts a limit on how much torque an engine can develop unless forced induction (a supercharger) is used. For an ultra-high-compression, normally aspirated engine (no means of mechanically boosting the intake charge), something a little over 1.65 ftlbs per cube (101 ft-lbs per liter) is about the best that is currently being achieved. With the aid of some type of efficient supercharger, the torque per cube can be doubled, tripled, or even quadrupled, but there are consequences, which I deal with in Chapter 5. Now we know what torque and horsepower are, so let us look at what it takes to make as much as possible of both these factors. Chart 1-1 shows all the elements that affect the engine’s final output. But everything has its limits, and before we go on to discuss how best to increase engine output, it serves us well to understand what limitations we may be faced with.

When Isn’t Bigger Better? You may have heard it said that an engine is nothing more than a simple air pump. Reduced to basics, it can be looked upon as an air pump, but the reality of producing horsepower means that it is hardly simple. A much better description is that it is a complex thermodynamic air pump. If high output is sought, the first aspect we need to pursue is the production of torque, closely followed by all the means possible to raise the RPM to where that torque occurs. Just how much torque an engine makes is governed by the weight of charge inhaled in one working cycle. We can increase that charge weight by making the engine bigger, by supercharging it, or both. At first thought, making the engine bigger by virtue of a bigger bore and stroke seem like a process that can be carried on without limits other than the sheer size of the final engine. Therefore, if you doubled the displacement of an engine, you may think that the output would double. Unfortunately, things don’t work out that way.

I stretched this Ford Windsor engine from its original 351 to 425 ci by means of a 0.6-inch stroke increase and a 0.060-inch overbore. To make the most of this capacity increase, it was necessary to change valve events and lift to more appropriate values.

I tried to dig up a big 5-inch aero-engine piston for this shot, but all I had at hand was a 4.185 small-block Chevy piston to compare with the 1.250inch-diameter piston from a gas-powered lawn trimmer. Big pistons may

look good, but you need to remember that the cylinders they reside in are far harder to fill than small cylinders. As the engine’s cylinder proportions are increased, it becomes harder to extract a proportional increase in power. Just after World War II, it was considered that, in normally aspirated form, 125 hp per cylinder was about the limit with the materials, oils, fuels, and technology then available. These days, better materials, modern fuels and oils, flow benches, and valvetrain dynamics have allowed the 125-hp figure to be exceeded by a big margin. Currently, a good V-8 Pro Stock “mountain motor” in the 825-or-so cubicinch (13.5 liters) range can put out about 200 hp per cylinder. But it hasn’t, nor will it ever, come easily. The reason for this is largely a question of proportions and geometry. It is better, if designing a given size engine from scratch, to have a lot of smaller cylinders rather than a few large ones. At first this seems less than logical, so let’s go through some clarification logic.

Here is the intake valve from a 1,050-cc-per-cylinder Pro Stock engine compared to the valve from a 23-cc-per-cylinder engine. Given all the same proportions, the small cylinder with its small valve breathes easier than the big one. Even if all the engine’s proportions remain unchanged, it is more difficult to fill a large cylinder than a proportionately identical but smaller cylinder. Here is an example to illustrate the situation: Assume we have a 180-ci (3-liter) one-cylinder engine as a starting point. Let us also assume that this engine has what is a fairly typical bore/stroke ratio of 1.2:1 (bore is 1.2 times the stroke) and the intake valve is 50 percent of the diameter of the bore. If this 180-inch engine has just one cylinder, the intake valve would be 3.252 inches in diameter, the valve’s

area would be 8.306 square inches, and its circumference would be 10.22 inches. For a cylinder head equipped with this valve to breathe effectively at the cylinder’s limiting RPM, it would need a minimum lift of about 0.980 inch (25 mm). If our 180-ci engine obtained its displacement from eight proportionally similar cylinders, the intake-valve diameter would be 1.626 inches. In this instance, the combined area of all eight intake valves is 16.612 square inches and the circumference is 40.88 inches. The lift needed to get nearmaximum flow is only about 0.487 inch. This means the eight small cylinders of 180 ci have approximately twice the breathing power of one big cylinder of 180 ci, even though the valve and bore proportions remain unchanged. Also, if we assume 5,000 feet per minute maximum piston speed, the maximum RPM of the one cylinder engine is 5,500, where it is 11,000 for the eight-cylinder engine. There are other factors involved, but from this you can see the reason it is harder to extract power from big cylinders rather than a greater number of small ones. You can also get an idea of how, ultimately, these and other limitations call a halt to the amount of power that can be had per cylinder.

The bigger the engine, the more important heads become. This CNCported head has the biggest valves possible. It also has thermal barrier coatings to keep the heat out of the intake charge in the port, and in the burning charge in the chamber. Since most of us build an engine using an existing block, rather than designing from scratch, understanding these limitations allows moves that help optimize what can be done within the constraints of a set number of cylinders. It is often said, and rightly so, that a successful engine is the sum

of a parts collection that is the right combination. What often defeats many engine builds is the fact they are not an orchestrated combination.

This Jon Kasse big-block Ford displaces a massive 900 ci (15.1 liters). The output from this 6-inch-stroke gas-burning behemoth is around the 1,700-hp mark, with torque just short of 1,600 ft-lbs.

CHAPTER 2

PRIMARY POINT INDUCTION

In this chapter we deal with atmospheric conditions and composition, and how to maximize the power that can be made from those prevailing atmospheric conditions. We are dealing with an air breathing heat engine here, so let’s start off with the air our engine has to work with. At what we call Standard Temperature and Pressure (STP) conditions, atmospheric pressure at sea level is 14.696 psi (1,013 millibars or, as measured with a mercury

barometer, 29.9213 inches of mercury). At 70 degrees F the density of air at 14.696 psi is 0.074887 pound per cubic foot. Assuming the air is dry, we find that, by volume, 21 percent of it is oxygen and 78 percent is nitrogen. The remaining 1 percent is composed of trace gases. But in most parts of the world, the air is far from dry; it has a water vapor content. Since water vapor does not support combustion, an even lesser amount of the whole is oxygen available for combustion. Assuming we are not racing under less-than-normal conditions, the atmospheric conditions just described are about the best we are likely to be able to see. At this point, things can only go downhill from here, unless we take steps to the contrary and that is what we investigate next.

Air Temperature Assuming there are no unwanted side effects, the lower the inducted air temperature, the more torque (and consequently more power) your engine will make. Underhood air temperatures can reach the top side of 180 degrees F without any difficulty. If you live in a climate like the hot arid conditions of the Southwest, those temperatures can reach 240-plus degrees F while the outside (ambient) temperature may only be 105 or so. Under these conditions, your engine could be giving up as much as 13 percent of its output everywhere in the RPM range. Since this brings about a reduction in torque, it also means that your gearing for best acceleration is also likely to be off by a similar amount. The result is a sizable increase in the 0–60 and quarter-mile times.

Picking up air from a cooler source produces more power by a direct increase in torque. An increase in torque throughout the RPM range has a significantly bigger impact on performance than just a top-end output increase. Simple Cold-Air Induction Most auto manufacturers pay only minimal attention to cold-air induction. They build induction systems to keep the air within a narrower temperature spectrum than the prevailing ambient temperatures. This makes it a little easier to meet emissions and, while this is good, what it does do is leave a lot on the table for hot rodders to exploit. In Chart 2-1, you can see just what a K&N cold air intake did for my GMC 4.8 truck, which I use mostly for towing. Just installing this cold-air kit produced, on a typical 75degree day, an average increase of a little more than 8 ft-lbs and 7 hp, with peak power going up by 13 hp. At the track a few days later, under similar before-and-after test conditions, the cold-air-equipped truck went 0.13 seconds and 2 mph faster in the quarter-mile.

Extreme Cold-Air Measures A cold-air inlet is a simple and effective means to achieving a little more output. Knowing it works, let’s apply the cold-air principle in a more extreme manner. Let’s start the ball rolling here with a very important point. Inducing cold air is always of value but, to make the most of cold-air induction, it is necessary to prevent heat from subsequently getting into the cooler charge so that it arrives at the intake valve as cool as possible. Doing this can achieve some very worthwhile results. Tests, shown here (in Charts 2-2 and 2-3) with a 5.0 Ford Mustang, are very revealing. Also, these tests demonstrate that heat can get into the intake charge by means not always quite so obvious. The 5.0 Mustang intake is a good example because it is one of the worst socalled high-performance intake systems on the planet. The reason being: It has an extremely high surface-area-to-volume ratio, and this makes it act like a heat sink that absorbs every BTU within 50 feet (well, not really, but you get the picture).

This K&N cold-air intake for a GMC/Chevy truck has the inlet end sealed from the rest of the engine compartment and picks up air from directly behind the headlight.

The pipe that leads from the K&N filter to the throttle body is of ABS plastic. It doesn’t conduct heat from the engine bay into the air passing through as easily as an aluminum tube would. However, an insulating wrap here would probably have helped. If you have ever watched a 5.0 Mustang owner who is intent on achieving the best quarter-mile time possible, you have noticed they are almost for sure packing ice on the intake before making a run. At the end of the run, the intake temperature often rises from a little above ice cold to more than 200 degrees F. In an effort to improve the stock manifold, the first move made for these tests was to pull it from the engine and make any simple mods to improve the airflow. But these “porting” modifications also cut intake charge temperatures. Several of the runners of the lower half of the 5.0’s fuel-injected intake have a severe “dog leg” just before entering the cylinder head port. By welding a little on the outside of the manifold, these dog legs can be straightened somewhat. After that’s done, I send the lower and upper half of the intake to be “extrude honed.” In this process, a highly viscous compound containing grit is forced through the runners. This produces a very smooth surface finish, and in so doing achieves two things: First, and most obvious, is that it smoothes out casting imperfections that may have snagged the airflow somewhat. Second, the smooth finish has far less surface area than a cast finish, so the charge picks up less heat from the walls of the intake.

2-1. Here are the curves as measured at the rear wheels. For this test the converter was in the locked mode and the transmission locked in third gear. Note that the torque gains are across the board, so this engine now drives as if it had slightly more displacement. This produced noticeably faster off-the-line acceleration.

Extra torque is always good when towing. For the quarter-mile, you can see from the chart how much farther ahead you would be at the end of the strip with the cold-air system in use.

2-2 and 2-3. Seen here are the results of a basic port job and extrude hone on the lower half of the 5.0 intake, and an extrude hone on the top half. Although these flow increases are significant at mid- to high-engine speeds, they are almost of no consequence at low speed because the stock intake is more than capable of dealing with the engine’s low-speed demand.

Final flow tests of the coated intake showed the 0.005-inch-thick internal coating produced no measurable reduction in airflow compared to the bare extrude hone finish. Charts 2-2 and 2-3 show the result this porting/extrude-hone exercise has on flow. This exercise produced positive power results, which are shown on page 15. After dyno testing with the ported and extrude-honed manifold, it was removed and subjected to an intensive thermal-management program. This involved applying a zirconium-oxide coating over the entire exterior and base surfaces of the manifold. Along with that, the inside of the runners had a 0.005-inch thick, thermal-barrier coating applied. To minimize heat passing from the hot oil splashing on the underside of the manifold, a splash tray was fabricated and the gap between this and the manifold itself was filled with structural insulating foam. The throttle body spacer was also modified to eliminate the heat passage.

Shooting zirconium oxide with a flash produces this glaring white image. Zirconium oxide also gets dirty really quick, so a thin coat of white lacquer is advisable here.

How did all this work out? Here the results about speak for themselves as Charts 2-4 and 2-5 shows. The chart’s baseline tests are for a system already equipped with a C&L cold-air intake system. The aluminum tube from the K&N filter to the mass airflow meter had the internal-diameter thermal barrier coated to cut heat transfer at this point. This system was initially worth about 8 hp, so the sum total gains from the porting/ extrude hone and the thermal management was substantial at 35 ft-lbs and 44 hp! In addition to track and dyno results, some temperature recordings with an infrared heat gun were also made. A typical 5.0 intake manifold tops 200 degrees F. But with the weather prevailing during these tests, it was more like low 190s at the start of a pass and mid to high 190s at the end. In our test vehicle’s case, the starting temperature typically was about 140, and it dropped to the mid 120s after one pass. It went as low as 105 after a second pass, which was done immediately after the first. For regular street driving, there is plenty of time to heat soak, but despite this, the temperatures rarely rose above about 145 degrees F.

2-4 and 2-5. The black curves show the rear wheel output with a C&L cold-air package consisting of a K&N filter located under the fender. This was our baseline. The blue curve is the output after porting/extrude-hone operation. Note the increase in low speed due not to flow increase but to

heat reduction. The red curve is for the extreme thermal-management system. This intake proved to be so effective that it got progressively cooler as back-to-back drag strip passes were made. Temperatures bottomed out at about 100 to 105 degrees F with the ambient at 75 to 80. From the forgoing, we can see that actively seeking the coldest source of air is a good move toward making more horsepower. There are, however, a few points you should be aware of to realize the best performance from any cold-air source. If the engine is fuel injected, then any possible air/fuel ratio error will more than likely be compensated for by the engine’s ECU. If the engine is, as is so often the case, a carbureted unit, then a significantly cooler intake charge almost certainly needs some carb calibrations to make the most of it. The first point is that denser air almost certainly needs a little more fuel. Usually a carb has a certain amount of self compensation, but extremely cold air may require an increase in fuel to restore the optimal mixture ratio. The second factor here is the carb’s fuelatomization function. For good ignition properties and initial flame propagation, a certain amount of fuel needs to reach the cylinder in the form of vapor. To compensate for reduced vaporization, it may be necessary to go with a booster design with greater atomization capabilities.

Ram Air: The Basics We have seen how thermal management can have a dramatic effect on power output. Now it’s time to look at another part of the intake air density equation: ram air. The big air scoops, as seen on cars such as Formula 1, Pro Stock drag cars, and the like, epitomize this means of power boosting. If all we had to do was look fast, then these scoops succeed in grand style. But going fast and winning races is about function, not style, so the question we have to ask here is: Do they work? Sure they do, but let us not overlook that they are also cold-air intakes. That said, let’s get down to analyzing the possible gains that may be had by using the forward motion of the vehicle to increase the induced-air density. In engineering terms, the increase in air pressure brought about by bringing a stream of air (moving in relation to an object) to a dead stop is known as the static pressure head. Two factors that affect this “ram air” pressure are relative air speed and air density. The ram air pressure goes up

in proportion to the air density but changes as the square of the speed. This means: double the speed, and the pressure goes up by a factor of four. In Chart 2-6, you can see that, at the speeds most of us travel, the potential gains from ram air are slim to minimal. To put things into prospective here, at 100 mph, the absolute best you could hope for is an increase in air density within the scoop of 1.2 percent. That is about the same as given by a 12-degree-F drop in intake temperature. If all this is starting to look like hard work for a small increase, you can rest assured; we have not reached the end of issues that are potential problems in making ram air work. The first is that the air is not brought to rest within the scoop, so the full effect of the vehicle’s speed is not realized. Remember, the scoop is being fed air at one end while the engine is sucking it out at the other. The scoop needs to slow the air after it has entered the scoop, thereby converting some of the kinetic energy into pressure energy. This is one of the reasons hood scoops on faster drag cars are bulbous in shape.

This hood scoop on Ricky Smith’s Pro Mod car dominates the hood but not without good cause. It is both the source of ram air and cold air. A scoop this size is only justified if the engine makes the horsepower to warrant it. Your average 800-hp drag racer needs something about 40 percent smaller than this.

2-6. In the formula, “Ad” is atmospheric density (0.0749 lb/cubic foot is used here) and “V” is speed in MPH. Note that it takes a speed of almost 170 mph to generate an increase in pressure of just 0.5 psi. At sea level, this represents an increase of just 3.4 percent. If the engine is carbureted, then another issue can raise a specter. If only the throttle body experiences the ram air pressure, the pressure difference that draws fuel from the float bowl to the booster is reduced and causes the mixture to lean out. To avoid this, the float bowl must see the same ramming pressure that the carb venturis see. Even though this may be done, it is not a complete fix because the ramming pressure is more effective on the float bowl than it is on the venturis. If the engine is fuel injected, then, as long as the manifold absolute pressure (MAP) sensor is doing its job, the mixture remains within working limits.

Practical Density Issues So much for theoretical issues involved with achieving high-induction densities. Now let us consider how best to apply what we have learned and how to avoid some of the most common pitfalls. In 1979, I wrote an air filter feature article for Popular Hot Rodding magazine that started by claiming that air filter technology was not widely understood, even among professional racers and car builders. As I write this book, it seems that the situation has improved and many have a better understanding of air filter technology, but they still have a long way to go,

especially among drag racers. Hopefully, what you read here gives you a far better understanding than you had before. So far, we have discussed the reasons for picking up cool air and possibly ramming it into the engine. All this is of little use if the air entering the engine is dirty and proceeds to carry in grit that will quickly wear out the rings and bores to the point where they don’t seal. The air obviously must be cleaned by an air filter.

Significant Pressures and Flow To determine where losses may occur in an air box/filtration system, it helps to analyze the box/case and the filter element itself as two separate but inter-related entities. To do its job effectively, a filtration system must perform in three areas. First, it must remove micron-sized grit particles from the air. Second, it must do this without incurring any significant restriction. And third, in the case of a carbureted engine, it must suppress booster buffeting. To see how all these sometimes conflicting requirements can be met, let’s start by investigating various filter elements and their effect on flow.

2-7. The simple manometer setup shown here allows a quick determination of the effectiveness of both a case and an element. For a high-performance engine, a difference in the legs of the U- tube of 1½ inches should be considered the maximum allowable. It may take some effort to get to this low a restriction, but big improvements can be made by selecting a highflow element and making obvious moves to improve case flow.

These tests show just how much variance there can be between a really good filter element and some inferior ones. I like to think that the publication of tests such as these had an influence on the market. Since the publication of this and similar tests, Uni-filter has ceased production, Fram introduced a copy of the K&N (the K&N must not have been as bad as they kept trying to tell me it was), and Amsoil replaced their dud with a very good, long-life synthetic-fiber element, which complements their excellent oil products. Tests conducted in Arizona in summer proved that taking time to improve element flow and cutting intake heat can be very worthwhile. Using a K&N element and cold air, via the reverse-facing hood scoop, cut the quarter-mile time of my emission-laden 305 Pontiac Trans Am shown here by almost 0.7 second, while speed improved by a solid 3½ mph. Filter Flow An air filter element’s flow capability is dependent on both its size and the characteristics of the material from which it is made. The physical size of the element is an easy assessment to make but the possible characteristics of the element are not. The chances are that the stock setup you have on your street machine is too small for the job. Inadequate flow in this area can be

attributed to the case, the size/type of filter element used, or both. You are usually correct if you assume “both.” If you prefer to establish beyond doubt just how much of a restriction the case and element may be, it’s not hard to find out. All that is needed here is to make a pressure tapping into the case and measure the depression with a sensitive pressure gauge or manometer, as shown in Chart 2-7. The technique for testing is simple. Find a reasonably steep hill and power it up in, say, second gear, from a relatively low speed. Have an assistant check the manometer readings at certain key RPM points such as 2,000, 3,000, 4,000, 5,000, and 6,000 rpm (more if the engine is a small high-revving unit). At each point, note the total inches of difference between the two legs of the manometer U-tube. When testing most stock cases with a new, regular paper filter installed, you find that an 18-inch manometer doesn’t suffice much past the mid range due to restriction. To determine the proportion of the restriction caused by the case and caused by the element, repeat the test with the pressure tapping in the case, but outside the element. Be aware that some cases are good and some are abysmal, while most paper filters are so-so. Remember, a manometer check on a typical factory 4-barrel carb setup (as described here) may not show whether a filter case is too close to the carb mouth or not. For the record, a typical case lid needs to be a minimum of 3 inches from the air horn or the mouth of the carb. If power is the principle criteria, all this testing begs the question: What is the maximum acceptable pressure drop within the case and filter assembly? In a perfect world, the answer is zero, but achieving that is nearly impossible. However, improving flow into the system for more power follows a strict law of diminishing returns. In practice, we find that if, at peak RPM, the measured pressure drop is no more than 1½ inches of water, then you can assume for all practical purposes your no-significant-loss goal has been met. At such a minimal restriction, the effect on output is barely measurable, even on a very high-tech dyno. The fact that a filter is present in the system can, if it meets the flow demand of the engine, be of benefit in terms of output. This can occur to the extent that the engine actually makes more power with the filter in place than without it. If a carb inlet is hanging out in a high-air-speed environment, we find that the boosters can be buffeted by the large-scale turbulence they are likely to experience.

Many years ago, I conducted some booster signal tests on a 1,600-cc Ford Kent engine powering a Formula Ford racer. This was done at Silverstone in England. The measurements indicated a 20-percent variance in booster signal. From this, I deduced that the mixture suffered rapid fluctuations to the tune of some 10 percent or more. Just installing a K&N filter and case for the Weber carb showed a speed increase on the longest straight of some 2 mph! I played on the fact that most racers would assume the filter slowed the car by telling them we were going to take the speed penalty (which of course did not exist) because we had and could only afford one engine for the year. They all must have believed it because no one else used a filter for several years to come! Element Selection It won’t come as much of a surprise that there is a substantial difference between the best and worst filter elements on the market. The amount of bogus information is also sizable, although not as bad in 2010 as it was in 1976 when I decided to investigate filter performance in depth. What I found out, and more to the point publishing it, almost cost me a career. Because a then-small and relatively unknown company absolutely blasted the entire opposition at that time, I had magazine editors ban my work because they figured I must be on the take, and big filter companies’ called and told me my measurements must be wrong. Fram’s PR guy even called and told me such critical tests should be done by an engineer, not a journalist. He was not too pleased when I told him I was an engineer—not a journalist! The irony here is that some 25 years later Fram introduced the Air-Hog filter element, which is a direct copy of the K&N! All this controversy is like stirring up a hornet’s nest. The work I have done to bring real-world numbers to performance enthusiasts, who want factual tech and not public-relations tech, has even embroiled me in law suits. I have written much on this subject and I feel that repeating it all here, to again prove a point, is unnecessary. If you want a stack of numbers, I suggest you read my book, How to Build Horsepower Volume 2: Carburetors and Intake Manifolds (published by CarTech, Inc.).

Because the K&N’s mode of function is so different from a paper filter, it exhibits a dramatically lower clog rate while still filtering to levels below those called for by auto manufacturers. All the filters tested here were the same size. The used K&N is the same part number as the new one. However, it has seen 80,000 miles (approximately 25,000 of it was spent on California forestry roads), which have a lot of dust and debris. Note that the used flow still exceeds that of one of the better paper filters on the market. What I am going to do here is to show some flow tests to highlight the fact that the K&N cotton/oil filter elements are just about the Rolls-Royce of filter elements (incidentally, Rolls-Royce mandates these for many heavy-

duty applications on its diesel engines). Since 1976, many companies have copied the K&N design, and this has led to the production of a number of really effective filters. However, before you take this information and go buy just any K&N look-a-like, I need to tell you there are K&N look-a-likes that are almost indistinguishable from the real thing that neither flow nor filter! My advice here is: Unless you are very sure of the filter you are buying, stick to the purchase of a K&N. It may cost more than a regular replacement paper element, but it is semi-permanent so you won’t be buying more filters down the road.

Element Clogging Other than flow and filtration capability, a filter element has to combat clogging. For a regular paper filter this is done by virtue of what might loosely be called excess area. The filtration takes place by arresting any particle larger than the holes in the paper. Unfortunately, each hole becomes, to a greater or lesser degree, clogged by the particles it stops. This results in less flow area for the air, thus increasing the resistance to flow. Here, the K&N has an advantage in that the engine-induced pulsating air causes the ends of the oil-covered cotton strands to vibrate. This causes what can best be described as a sweeping of the air as it navigates its way through the multi-layered element. This action tends to collect the dirt on the cotton strands themselves, not in the voids between them. This mode of function leaves us with an element that can hold huge quantities of dirt before flow starts to degrade by any real amount. This is why most off-road race vehicles use K&N (or functional copies of K&N) elements and not paper ones, which degrade from the moment of start-up. The Filter Case Now let us consider the case, or what often serves as the case. Look through any performance magazine and you find many ads for cold-air kits, be it for a Honda, a full-size late-model fuel-injected Ford, Chevy, or Chrysler truck. What isn’t so well catered to, short of a monster hood scoop, are V-8 engines often built by hot rodders employing a single 4-barrel carb. The good news here is that Air Inlet Systems in Hamilton, Canada, specializes in cold-air filter cases for just such applications (visit www.ramairbox.com to see many highly functional examples). This is a

company that truly has a handle on filtered cold/ram air. Some tests run on a friend’s car showed an ET change from 11.51 to 11.36, and an mph increase from 118.52 to 120.11 in before-and-after tests run within 15 minutes of each other. Using standard quarter-mile horsepower equations, this indicates the Air Inlet Systems unit was worth about 18 hp on a nominal 450-hp motor.

Zero Power Loss Using K&N Filtration and Sizing

B

uilding a zero-loss filtration system is not difficult. It’s just a question of knowing how much horsepower each square inch of installed area (diameter × height) will support. For normal street applications, my own testing has established that a round K&N supports a solid 7-hp per square inch with no measurable power loss. This compares to about 5.6 for the very best paper filter and about 5 hp for a foam filter before it is oiled (and that can drop to as low as 3.5 hp when oiled). If you intend to use your vehicle in off-road conditions, then the filter needs more initial flow, so that there is more in reserve as the filter loads with dirt. Under these conditions, your K&N filter needs to have 1 square inch of area for every 4 to 4.5 hp the engine makes. Stick to these ratings and the filter element will be your engine’s best friend, not the power-robbing component it is often seen as.

Temperature Reduction—Adding Cool Fuel to Cool Air

I

t is entirely practical, after the air has been drawn in and kept near ambient temperature by any intake thermal insulation that has been used, to further cool the charge. Normally, there is a considerable reduction in intake temperature due to the evaporation of some of the

fuel as it moves through the carb and on down the runners to the cylinders. This normally accounts for a 30- to 50-degree-F drop in temperature on a typical race engine on an 80-degree day. But when the fuel evaporates and cools the charge, the evaporated fuel takes up more room; so much of the benefit of the evaporative cooling is lost. One aspect, however, can be a complete bonus. If the fuel is precooled before it enters the carb, the fuel’s reduced temperature can cool the carb itself and further cool the entire intake charge. This is normally accomplished by means of a “cool can.” The fuel flows through this can, which has within it a coiled copper tube. The can itself contains a mixture of dry ice and alcohol or dry ice and straight anti-freeze in it. When the fuel is cooled this way, it can result in a significant chargetemperature reduction. However, the resulting reduced evaporation may call for a booster that atomizes the fuel more finely.

CHAPTER 3

FUEL DELIVERY SYSTEMS

3-1. To achieve good results, both the mixture ratio and the fuel mixture quality need to fall within close limits. What is actually needed changes as engine loads and driving conditions change. So far, I have addressed ways and means of maximizing the air density delivered to the engine’s induction system. However, all this is to no avail unless the fuel is mixed with the air in a well-defined manner and to equally well-defined proportions.

From Chart 3-1, it might look like we are out of step with the sequence of events that lead to power. We are at the “Optimize Mixture” step and have not yet looked at the factors that apparently precede this. The reality here is that as long as the mixture is of the right proportions and is suitably finely atomized in a proportion to aid combustion initiation, it is of little consequence how it got that way prior to arrival in the cylinder. In practice we find that, however good the mixture of fuel droplets and air may be as it is discharged from the carb booster venturi or injector, things almost always go downhill from there. This means that not only do we have to do a good job of atomizing the fuel at the point it enters the inlet tract, but we also have to employ measures that help maintain good mixture qualities right through the inlet tract and on into the cylinder. At this point, it is clear we have to deal with three issues: mixture ratio, exhaust pollutants, and mixture quality. Let’s start with the mixture ratio.

Mixture Ratio Getting the mixture right is important; it can cost a lot of power if it is not. The margins within which we must work are relatively small, as can be seen in Illustration 3-2. In this illustration, we are holding the fuel volume constant and changing the air volume to show what is physically being dealt with. The fuel represents a relatively small volume compared with the volume of the air involved. For a rich mixture, there is more fuel than can be burned completely by the oxygen available. This means that after combustion there is unburned fuel contained in the exhaust (hydrocarbon emissions). When there is 10- to 15-percent excess fuel, maximum power is developed. For a stoichiometric, or chemically correct, mixture there is exactly the right amount of fuel to completely utilize the air; so, theoretically, 100 percent of the fuel is burned and 100 percent of the oxygen contained in the air is used to do so. When the mixture is lean, there is surplus oxygen remaining in the charge after all the fuel has been burned. It is under lean-burn conditions that the best fuel economy is developed.

3-2. Seen here are the relative amounts of fuel and air in the proportions we have to deal with. Note the great difference between the amount of fuel and the amount of air. This is our first indication that calibration accuracy is critical.

3-3. There are some important factors displayed in this graph. Be sure to read the nearby text for the relevant issues at stake here. There is a somewhat famous statement made by someone whose name no one seems to remember that goes like this: “A carburetor is a wonderfully ingenious device that delivers the incorrect mixture at all engine speeds.” Well, that statement is sure to catch attention, but it is totally wrong. By the middle of the last century a good carb, properly calibrated, could deliver mixture accuracy as well as any modern electronic fuel injection setup. Now let us consider the effects that a change in mixture has on wideopen throttle (WOT) output. Chart 3-3 shows just what can be expected in the way of a change in power in relation to a change in mixture. The width of the green column is our target and, as can be seen, this is not a very wide target to be shooting for. That is why it is so important to get the mixture right. As the chart shows, output drops off far faster on the lean side of the curve than on the rich side. Therefore, the calibration should err slightly on the rich side because any power loss is smaller than the same error on the lean side. The target zone (green column) for a gasoline-fueled engine is typically between 13.2:1 and 12.5:1. Anywhere between these two ratios, the

change in power is not usually measurable. However, if the engine is to be used for long-distance racing, fuel economy (to cut the number of pit stops) also becomes a requirement. Under these conditions, running the engine slightly on the lean side of the peak power point (13.5:1) is a worthwhile strategy. If the engine concerned has good fuel distribution between cylinders, the optimum power ratio is right on 13:1.

3-4. Other than oxygen (O), which is left over under lean-burn conditions, what we see here are the basic exhaust pollutants and how they typically change with the mixture ratio. What is not shown here are oxides of nitrogen (NOx), which are a product of high pressures and temperatures. The blue zone of the graph represents the air-to-fuel ratios we should target for cruising with fuel economy. Here, we are obviously looking to go the farthest distance possible on the least amount of fuel. To do this we run the engine lean; that is, less fuel than the air is capable of burning. To a point, the leaner we can run the engine, the more fuel efficient it is; but as it is leaned out, the harder it becomes to ignite the charge. At this stage, spark energy and temperature play a more important role. A good

ignition system (and most modern ones are) effectively lights off a mixture as lean as 17:1, but leaning beyond 17:1 does cause a lean-mix misfire. With suitable design of all the pertinent factors and a very high intensity spark (large and of high temperature), I have successfully run mixtures as lean as 22:1, and some research establishments are reporting 30:1. Running mixtures as lean as this produces some really good fuel-efficiency numbers, but getting there is not a simple job by any means.

Exhaust Pollutants Although it is not nearly as relevant to performance as overall calibration, we should, in this day and age, consider the effects of our tuning on exhaust pollutants. If you are intent on running an engine without a catalytic converter, then you sort of owe it to the rest of the world to do your best to save the planet. Alone, your effort won’t be of much consequence. But if we look at everybody who is likely to read this book, making sure the exhaust is as clean as possible is relevant. Chart 3-4 shows what we are dealing with here. What this chart shows is that the exhaust is essentially cleaner as the mixture becomes leaner. This means the time spent setting up that idle, so that it runs as lean as possible, is a worthwhile step toward not only saving fuel but cutting pollution. The same goes for the cruise calibration. Keeping things on the lean side not only helps cut pollution and save on the fuel bill, but your engine also lasts considerably longer in the bore and ring department. Running mixtures from 14.7:1 on up compared to, say, 13:1, can increase bore life by as much as 400 percent. When that throttle goes wide open and maximum power mixture hits 13:1 we are going to see quite an amount of unburned hydrocarbons, along with 6 percent of the exhaust consisting of carbon monoxide. CO, as you should be aware, is poisonous. Since the duty cycle of most performance engines is 99-percent idle and cruise, fuel delivery calibrations at these operating positions are far more important than what happens at WOT.

Mixture Quality

I don’t address the important subject of mixture quality in too much detail here, but I do bring it up where relevant in subsequent chapters. Suffice it to say that fuel will not burn in a liquid state. That means fuel rivulets within the induction system are a power detraction. The mixture needs to be finely atomized so that the bulk of the fuel arrives in the cylinder as suitably small droplets. Just how small they need to be depends on the temperatures involved. For the ignition system to effectively light off the induced charge, there needs to be at least 15 percent of the fuel vaporized. However, if too much is vaporized within the intake manifold, the greater volume taken up cuts the amount of air drawn in. This means that too much vaporization reduces the engine’s volumetric efficiency (breathing efficiency), and that is not good for power.

With a carbureted engine, good mixture preparation starts at the booster and the air corrector. Understanding how each of these functions and their interactions can be worth more than just an average race-winning margin.

When this totally stock, injected ZO6 346-ci Corvette engine was tested on the MotorTec Magazine dyno, it cranked out 412 hp and 401 ft-lbs. All this was done while meeting stringent emissions regulations. A high-tech intake manifold and fuel injection were key elements in meeting these impressive results. Getting the right balance of droplet size at a given temperature is not so much an issue with a fuel-injected engine because the typical injection pressure of 45 psi and the (usually) close proximity of the injectors to the intake valve help ensure positive results. That situation does not necessarily exist for a carbureted engine. Getting the mixture quality to be as near optimal as possible involves paying a lot of attention to the fuel delivery at the carb’s booster venturi (auxiliary venturi). After this, you need to pay attention to port runner shape and finish to minimize the negative effects of fuel drop out.

Carb Function Basics Virtually all carbs function by creating a reduced air pressure over some kind of metering jet connected to a fuel bowl. There is a type of carb known as a constant-vacuum carb. The most popular and well known of this type is

the SU carb, which has been used on British cars for nearly half a century. Although they could be made to meter fuel extremely accurately (as good as any fuel injection), they were not popular beyond the shores of England. That being the case, I focus more on the type of carb known as a fixed-jet carb. This employs a venturi to create the depression that draws the fuel into the intake system.

3-5. The style of Holley carb seen here has won more races than all other carbs put together. That being the case, I use it as our primary example on which to base the explanation of the basic principles of carburetor function.

The Quadra Jet, used by so many of GM’s V-8-powered cars right through to the early 1980s, can be viewed as a hybrid carb. The very small primary side has a fixed-jet mode of operation while the secondary is a version of a constant-vacuum carb.

These side-draft Webers epitomize the fixed-jet genre of carburetors. Although outwardly complex, a basic knowledge of carb design reveals they are far less complex in nature than might be previously thought. Illustration 3-6 clearly shows how the increase in velocity within the venturi causes a drop in air pressure at that point, which draws the fluid from the reservoir up the tube. With a little increase in speed, this example would start discharging the fluid from the reservoir. Of course in real life there is only one connection between the reservoir and the venturi: the one at the minor diameter. Now, if you are observant, you may have spotted a problem here. Because the fuel needs to be a certain distance below the discharge hole (known as the spill height) quite a lot of air must flow before any fuel is drawn into the intake tract. So this system won’t work for an engine at idle. In addition to this, the system won’t deliver twice as much fuel when the airflow increases.

This cutaway of a Barry Grant Demon carb shows the main jet delivery system in red. Although all this may look complex, understanding the basic principles involved unravels the complexity to a great degree.

3-6. As speed increases at the venturi’s minor diameter, so the air pressure drops and, in so doing, it sucks fluid up from the reservoir below. The height of the fluid in the tubes represents the amount of suction and where it is occurring.

Seen here is the application of the venturi principle to something a little nearer a real-life carb. Although this looks simple, the dynamics of increasing flow through a venturi and the spill height are going to make the situation more complex if any sort of metering accuracy is to be achieved.

The function of the air-corrector jet is illustrated here. As the draw on the main jet increases due to increasing venturi action, an increasing amount of the signal is bled off by the air corrector. Now a mixture of air and fuel is discharged from the tube in the venturi. This is commonly known as an emulsion.

The fix for this is an ingenious little jet known as an air-correction jet. Just to avoid any confusion, here it is (on page 27) on a Holley carb. It is known as a high-speed bleed, when referring to the main jet circuit. The aircorrection jet has little effect on the mixture at low engine speeds, but as RPM increases it has an increasing effect. By suitable sizing (in relation to the main jet), a fuel curve much more closely approaching a straight line can be produced from low to high RPM. With the addition of an air-correction jet, it might look like our calibration characteristics are pretty much sorted out. Unfortunately, that is still far from the case. An engine does not draw in air like a vacuum cleaner. We find that as the intake and exhaust go through periods of breathing effectively, and then not effectively by virtue of system resonance, the pull on the jets produced by the intake airflow vary in a seemingly near-random manner. What our simple carb needs at this point is a means of recognizing this and compensating accordingly. To manage the mixture calibration under what may be rapidly changing volumetric efficiency conditions (as RPM increases), a fixed-jet carb utilizes what is known as an emulsion tube. Although this may take on many different forms, a quick study of any example shows they all have a common mode of function. However, deciphering just how an emulsion tube and jet system may function just by looking at it is not so easy.

The brass jets on either side of the booster legs are air-correction jets. The outer jets are for the idle/transition circuit and the inner ones are for the main jet circuit.

This is the main-jet (far right), emulsion-tube (center), and air-corrector (left) assembly from a Weber carb. Although appearances may vary somewhat, this typifies the layout used for a fixed-jet carb. The carrier (far left) facilitates fitment into the carb body. In Illustration 3-7 we have an emulsion tube in the well it resides in. The well receives fuel from the carb float bowl via the main jet located at the base of the well. This fuel surrounds the emulsion tube. At the top of the emulsion tube, we have the air-corrector jet. The hole pattern in the emulsion tube modifies the air drawn in to the air corrector jet. If there were no holes in the emulsion tube, it would remove the effect of the air corrector from the system. However, what happens here is the air drawn in from the aircorrector jet pushes the fuel level down in the emulsion tube. The air then bubbles out of the emulsion tube into the fuel in the emulsion well. Now, instead of there being neat fuel in the emulsion well, the fuel surrounding the emulsion tube holes is a mixture of fuel and air. In other words, it is an emulsion; and it is delivered to the venturi as part fuel and part air, rather than straight fuel. By modifying the ratio of fuel and air composition of the emulsion, we can modify the overall fuel/air ratio delivered by the carb.

3-7. The basic function of the emulsion tube is to bleed air from the aircorrector jet down into the volume beneath the air corrector and on out through the holes in the emulsion-tube wall. By suitable placement of the holes in the emulsion-tube wall the fuel curve can be modified to suit the volumetric-efficiency curve of the engine.

Here, indicated by the arrows, is what goes for an emulsion tube on a Holley carb. Most Holley carbs are used on a plenum-style intake, rather than an independent runner (IR) system. When induction pulses are smoothed, as is the case when all eight cylinders of a V-8 draw from a common source, we find the need for an emulsion tube and the degree to which it affects the mixture is decreased. In this photo, we see a typical emulsion-hole pattern for a typical street engine (left) and the adjustable jetted system (right) used for fine tuning the fuel curve on a serious highperformance engine sporting a race or race-like cam. To calibrate the emulsion tube and well operation, the first step is to know where in the RPM range the mixture goes rich or lean. The next step is to understand that the top of the emulsion tube affects the bottom of the RPM range, and the bottom of the emulsion tube affects the top of the RPM range. Looking at the emulsion tube and relating its length to the RPM range, we can now say: Where there are holes in the tube, the mixture is leaned out; where there are no holes, it will be enriched.

Contrary to popular belief, the main jets of a Holley carb (shown here with jet extensions) are the principal means for calibrating the high-speed cruise mixture, not the WOT mixture. The power-valve calibration is actually the primary influence at WOT. An example should make all this clear: Let us say we have arbitrarily selected an emulsion tube pattern and find it delivers the right curve, except at the lowest of engine speeds where it is too rich. Remembering the top end of the emulsion tube affects the low speed, we fix this problem by selecting an emulsion tube with more holes at the top. This bleeds more of the aircorrector jet’s air into the emulsion tube at this point, and thus leans-out the mixture at the low end of the RPM range while having minimal effect on the mid and top ends. If the mixture is correct at low and high speeds but rich in the middle, then we add holes in the middle of the emulsion tube. Armed with this information, you should now be able to rectify any fuel-curve issues a fixed-jet carb may be having.

3-8. The power valve opens when manifold vacuum drops below a certain predetermined level. This allows fuel from the float bowl to enter the emulsion well through both the main jet and the power valve restriction channel (PVRC). If the fuel calibrations are done with both optimal power and optimal cruise in mind, then the WOT calibrations should be done with the PVRC, not the main jet. The main jet should be sized for the best fuel economy under high-speed cruise conditions. In the right-hand illustration the power valve is closed. In the inset enlargement (left), it’s open and the fuel flow route can be seen.

WOT Calibrations

With most fixed-jet carbs, the mixture calibration is moved up or down the scale by the sizing of the main jet. Install a bigger main jet and the mixture becomes richer throughout the RPM range. Install a smaller one and the opposite happens. Sounds simple enough, but the main jet calibration still must deal with the mixture requirements that, though fairly wide open, still fall short of WOT. With an independent runner (IR) system that employs a Weber or similar carb, we find that the undamped induction pulses at WOT cause a richening of the mixture. As the throttle is closed part way, the pulses are damped and this conveniently leans-out the mixture to something in the order of what is required for lean (or leaner) cruise. When a carb such as a Holley is used on a plenum-style intake, induction pulse enrichment is almost nonexistent. This means enrichment from a cruise mixture of, say, 15:1 to a full power mixture of 13:1 has to be handled differently. This is typically done by means of a so-called power valve circuit. A diaphragm in the power valve senses when the intake manifold vacuum drops below a certain value and, at that point, opens up an additional jet, which then supplies more fuel to the booster venturi. Illustration 3-8 shows the interaction between all the principle components in a Holley carb when going to WOT operation.

Here is the metering plate of a Holley carb destined for a circle-track race engine. Note the calibration jets (arrow) in the PVRC holes.

The Boosters

We have worked our way through the system and now it is time to take a look at the critical job done by the boosters. First: Why the term booster? This is so called because its function is to amplify the pressure drop or signal generated by the airflow through the main venturi. Depending on its design, a booster can amplify the main venturi signal from 80 percent to as much as 480 percent. By amplifying this signal, you attain the means to achieve both finer atomization and more accurate calibration. Without the aid of a functional booster, the carb would have to utilize a much smaller main venturi, so that sufficient signal would be generated at the low end of the RPM range. But a small main venturi would then limit the top-end output due to insufficient total airflow. As you can see, the importance of a booster being compatible with the operating needs and RPM range of the engine is starting to shape up as an important issue.

Unlike a Holley carb, a set of Webers on an IR manifold needs no power valve for WOT fuel enrichment. At WOT, the increased amplitude of the induction pulses causes the mixture to become enriched over that seen at part throttle. The first step toward knowing what to expect of a booster is to understand how it works. Illustration 3-9 shows the essentials. The principle is simple. Air is drawn through the main venturi due to a pressure drop at the exit. The velocity of the airflow at the minor diameter of the main venturi produces the characteristic drop in pressure at that point. Because the end of

the booster is located at the main venturi minor diameter, the booster sees a greater suction drawing air through it. This results in a higher velocity at the minor diameter of the booster, which produces a greater depression (signal) than seen at the minor diameter of the main venturi. As with most aspects of engine performance, more is not always better. For best performance at WOT, a carb booster needs to help produce the ideally sized fuel droplets for combustion. If the droplets are too large, combustion efficiency suffers. If they are too small, the fuel evaporates in the intake system too soon and reduces volumetric efficiency. For part throttle, where the best fuel efficiency is sought, it is best to have the fuel completely vaporized by the time the piston is close to the top of the compression stroke and the spark is about to fire.

Attention to idle and intermediate circuits is shown in this 1,250-cfm carb from AED that adds to the top-end output and is usable on engines as small as 350 inches.

3-9. The depression at P1 produces a greater depression at P2. Because the end of the booster is located at P2, it produces an even greater depression at P3.

Seen here in order of increasing gain and fuel atomization are Holley’s principal booster designs.

A vehicle is virtually undrivable without effective idle and transition (known in Europe as progression) circuits. This cross section shows how these circuits are laid out in a 4150-series Holley, but the principles apply to most other carb styles. Engine vacuum virtually drives the idle system. Fuel is drawn from the float bowl through the idle jet and is emulsified with air drawn from the idle air corrector and from the transition slot when the edge of the butterfly is below the transition slot. The amount of fuel emulsion going into the engine, and hence the idle mixture, is controlled by the idle mixture adjustment screw. As the throttle is opened, more air enters the engine. To meet the needs of increased fuel flow, we see that the transition slot changes its mode of function. As the throttle opens and the edge of the butterfly moves to above the slot, it stops metering air into the system and meters additional fuel instead.

Idle and Transition Circuits Up to this point, we have talked about the production of maximum power from our engines and what may be needed to cruise fuel efficiently at relatively high speed. As important as that is, a high-performance street engine spends as much time at idle and transitioning to low-speed cruise

(just off idle) as it does at, say, 60 to 75 mph. The bottom line is that the term drivability is almost all about getting the idle and transition circuits calibrated right on.

Acceleration Enrichment Let’s talk fuel vaporization for a moment. When an engine is at idle or cruise, the intake manifold is at its highest vacuum situation. At typical temperatures under these conditions, almost all of the fuel vaporizes. If the throttle is now opened, the absolute manifold intake pressure quickly rises or, in other words, the manifold vacuum decreases. Whatever fuel was vaporized in the air now rapidly condenses on the walls of the manifold. This means the mixture being pulled in at that moment is really too lean to burn. The result is a total engine cut-out. To have the engine perform as intended under the transient conditions of rapid throttle opening, we need to force-feed the system with extra fuel. With an injection accelerator and pump, the fuel delivery system has a means of adding extra fuel as the throttle blades are opened. With a carbureted system, this function is achieved by means of an accelerator pump. The idea is to have the accelerator pump (or the equivalent action for a fuel-injection system) rapidly pump in just enough additional fuel so as to cover the hole in the fuel curve during the opening of the throttle.

Here is a simplified layout of a carburetor’s accelerator pump system. We can see that rapid opening of the throttle actuates the piston, which in turn pumps in extra fuel just above the booster.

Like this Holley-style Braswell carb, most carbs use a diaphragm accelerator pump design rather than a piston.

A quick means of determining this is to have too much fuel as a starting point. This starting point can be observed as black smoke in the exhaust as the throttle is rapidly opened from a closed position. From here, it is just a question of cutting the amount of fuel pumped in until the engine just stumbles, and then taking a step or two back. As for calibration methods, we find that the amount of pump action is set by a cam or a spring-loaded rod on the throttle arm. This usually determines the rate and length of time for which additional fuel is injected. Other means of calibration involve sizing the pump and pump jet at the discharge point and a calibrated spill-back valve, which dumps a portion of the pump discharge back to the float bowl.

Carbs—How Big? With a fixed-jet carb, if the venturis presented to an engine are too big, we find that the air speeds through the main and booster venturis are too low to effectively atomize the fuel. What this tells us is that we need to size a carb so it functions perfectly at the lower RPM, while delivering the airflow requirements for the sought-after top-end output. There have been many methods employed to produce as wide an operating span as possible here. The Quadra Jet and the Edelbrock carb employ a small fixed-jet design on the two primary barrels of the four, with an air valve that acts like a constant vacuum carb on the secondary side. This mode of operation can be likened to “size-on-demand” and as such can be made to work well. For a Holleystyle carb, the answer for good all-around street performance is the vacuum secondary carb.

This 2x4 setup from Edelbrock looks like a lot of carburetion for a typical 350-ci small-block Chevy, but this proves not to be the case. First, a twoplane intake is used. This means that each cylinder sees only one of the two carbs. Also, the air-on-demand secondaries ensure sufficient velocity through the primaries until the engine needs the secondary flow. The first step toward installing an optimally sized carb is to make a preliminary selection based on the engine’s displacement. The next step is to modify this result by factoring in relevant engine spec details, such as the heads and cam used. For the initial calculation step, we need to determine the amount of CFM the engine is likely to inhale if it is able to breathe at 100-percent efficiency. To do this, we first multiply the displacement (CID) by the anticipated RPM the engine is likely to turn to. It is important to be realistic when making the RPM estimation. Start by estimating where peak power is likely to occur and then add 500 rpm to allow for over-speed. Because we’re dealing with a four-cycle engine, which has an induction stroke every other revolution, we need to divide the CID × RPM result by 2. The resulting number tells us the cubic inches of air the engine displaces per minute. To change that to cubic feet, divide by 1,728.

3-10. Assuming the compression ratio and the exhaust are appropriate for the engine, the heads and cam duration remain as the most influential factors in selecting the right-sized carb. To obtain the correction factor (CF) for a particular application, first choose a curve relating to the cylinder head spec being used (from the list below). Next, locate the 0.050 cam duration figure that is to be used (from the choices along the bottom). Then, go straight up the graph until an intersection is made with the previously chosen curve. Now go to the left to find the CF to be used in the formula. Black = stock OE heads of pre-1990 design. Magenta = pocket-ported pre-1990 heads or stock Vortec/aftermarket heads. Blue = street-ported heads. Green = race-ported conventional heads. Orange = top-of-the-line race-ported heads such as those used by pro racers. Red = super race heads such as Pro Stock and Nextel Cup heads. An example goes like this: Cylinder heads, street ported, use the blue line. Cam: 240 degrees at 0.050. Follow the 240 line up the graph until it intersects with the blue curve, then go across to the left. The left-hand scale indicates that the correction factor is 1.04. Slot this number into the formula, and you have the carb CFM required for the job.

The King Demon from Barry Grant Carburetors is a Dominator-sized carb with replaceable main venturis. It can be sized from 800 up to about 1,300 cfm. On this version, the secondaries are vacuum operated. This makes the carb a good candidate for a street big-block engine in the 460-plus-ci range. Our calculation so far assumes the engine has a 100-percent breathing efficiency. For a race engine, where the exhaust scavenging is a factor, the volumetric efficiency can exceed 100 percent by quite a big margin. For instance, a well-built race 350, with no regulatory race restrictions placed on it, can reach about 115-percent volumetric efficiency. This means that the engine, as far as the carb is concerned, looks as if it is 400 ci. At the other end of the range, we find that an absolutely stock street engine may have a volumetric efficiency of only about 80 percent at best. This means that an engine of the same 350-inch displacement appears, from the carb’s point of reference, to be only about 300 inches. Our carburetor selection needs to take this into account. The airflow required by an engine depends primarily on the cam, and the breathing capability of the heads.

Looking for performance and mileage? The 550 Edelbrock carb (top) is dedicated to producing good mileage. On my test vehicles, the 800-cfm Edelbrock carb (bottom) delivered more torque as well as more mileage than the stock Q-Jet it replaced.

Carbs can successfully be used with a supercharger as a suck-through or, as seen here, a blow-through installation. In either case some special reworking has to be done to the carb to make it compatible. For a blowthrough installation, I recommend dealing with The Carb Shop in California.

Assuming that the compression ratio and exhaust system are appropriate for the engine, the heads and cam are left as the most influential components on carb size selection. As cams get longer, so the engine’s volumetric efficiency improves. The volumetric efficiency also improves as the cylinder head’s flow capability improves. Chart 3-10 gives a correction factor (CF) to take into account cam duration and cylinder head flow. Using this correction factor, we’re now in a position to come up with a simple formula that gives a good prediction for required carb CFM:

The carb sizing example we have just gone through is a little on the conservative side, as it makes no allowance for the fact that a tricked-out carb with high-gain boosters can successfully use greater CFM. This allows a little more power to be developed without sacrifice in the lower-RPM range. Going this route does mean you have to know your carbs or work with a carb specialist such as AED or the Carb Shop (to mention but two of the many out there).

Carbs Versus Fuel Injection Since about 1985, people have commented on the fact that I do not do enough with fuel injection, and that maybe I should get up to speed in that department. Just for the record, I have been building fuel-injected engines since the late 1960s. But the situation is such that due to my skills at sorting out carburetion problems both in basic design of such and calibration, I keep getting called back into the carburetion camp.

Here is the fuel-injected system I was using on one of my 5.0 Mustangs at the time of this writing. I am just about to install another cold-air system to evaluate on the chassis dyno.

This Mass-Flo injection system is a prime example of a simple-to-set-up system. It is probably simpler to calibrate than a carburetor. My experience with this type of system on engines in the 500-hp range is that they do a good job and can show better mileage than a typical Holley or similar carb. This is not all bad. When the use of electronic fuel injection began to take hold in the early 1980s, it was popularly predicted that carburetors would be dead in just a few years. Well, here we are and the carb business is still strong, and carbs for use on hot-rodded V-8s still dominate the scene. Why? Because they are cost effective, easy to understand (relatively speaking), and they flat-out work when correctly sized and calibrated. But does this mean that I am solely focused on carbureted induction systems? No, I actually prefer fuel injection, and the good news here is that systems

are becoming more user friendly as time goes by. Also the price is coming down!

Here is a prime example of fuel injection being put to good use. This intercooled, Magnuson supercharged 6-liter LS6 Chevy had to meet emissions as well as make big numbers on the dyno. It did, and the injection system played a vital role in achieving that goal.

I liked the results I saw from the 650-cfm Mass-Flo systems I’ve seen on the dyno. This prompted me to rework the intake manifold for maximum airflow and to make up a new mass-flow meter to throttle-body housing as seen here. This, in conjunction with a high-flow race mass-flow meter, resulted in a little more than 1,000 cfm.

Because this Ford three-valve 4.6-liter Mustang is stock-equipped with a very flexible and easy-to-retune injection system, adding a ProCharger turbine supercharger as seen here becomes a much simpler calibration task. This vehicle can still meet emissions while delivering some 460 hp at the rear wheels. To date I have run systems from (and can recommend) Holley, Fast, Mass-Flo, Accel, and a few others. In addition, I have also used factory systems that have been re-programmed. Basically, the difference between calibrating carburetors and fuel injection is: Carbs need to be persuaded to deliver what you think the engine wants, and with fuel injection you tell the computer what you think the engine wants. It is much easier to know where things are at in the second instance!

This is my intercooled Magnuson-supercharged GMC tow truck. Without the versatility of the fuel injection, building this emissions-legal engine would have been extremely difficult. The stock injection was used in this instance with no mods, except bigger injectors, and reprogramming of the stock ECU.

If you plan to run to a setup like this, then Blower Drive Services (BDS) can supply all the needs to install the supercharger on your engine and the fuel-injection system. I sat in on the dyno testing of this 540-inch blown big-block Chevy. It made torque and power by the spade load, but that was expected. What is really impressive is just how smooth it idles and drives just off idle. This is what constitutes a totally civilized performance engine.

CHAPTER 4

INTAKE MANIFOLDS In Chapter 3, we looked at what was needed in the way of mixture quality and air/fuel ratios. Now our focus is the design of the intake manifold. Here, I discuss what it takes to minimize flow restrictions and optimize intake pressure waves. Many of you are reading this book to get some idea of how much extra power you are likely to see from a manifold change. It is reasonable to expect that an aftermarket performance intake manifold actually can deliver a cost-effective amount of extra power. Unfortunately, this is not always the case. I have tested intake manifolds that gave really mediocre results. In the mid 1980s, I did a big intake manifold test for an article in Hot Rod magazine. From 15 different intake manifolds, only about four of them produced what I call worthwhile results. Two or three produced barely more than the stock intake, and one actually reduced the 325-hp test engine by a whopping 56 hp. To make sure it was not a problem just with my test engine, I sent it off for another race shop to test. Same result! The moral here is: You would think that if you have two manifolds that appear the same at critical points (such as runner junctions to a plenum, runner lengths, etc.), they should produce similar results, but that is not always the case. Why? Because not enough consideration was taken to understand what is needed.

Lack of suitable technology application when designing an intake manifold can be a major stumbling block toward achieving good torque, power, and economy. In this chapter, we look at what it takes to determine an effective design.

This is the intake from my 2002 Dodge Intrepid Cup Car. To feed an engine with a 750-hp potential, the manifold runners have to flow well and have good port velocity. Also the entry to each runner has to be taken care of for airflow, good wet flow, and fuel distribution. Photo 4-1 shows four intakes for a 2-liter non-VTec Honda. One might expect the intake that had the highest average flow was best. In this instance, it was the second intake from the top. However, it showed a minimal gain in top end (about 3 hp) while costing twice that much in the 3,000- to 4,500-rpm range, which is used so much more often. To make any plenum-style intake work, we have to consider the size and efficiency of whatever feeds the plenum in the first place. Next, the entry into the runner has to be optimized by suitably reshaping it. After that, the diameter and length of the runner itself has to be determined for best results. Fail at any one of these factors and the results suffer. Still using this Honda as an example, let’s start with how the plenum is fed. The object of the exercise here is to supply the plenum with as much air as needed with the minimum restriction. The obvious solution is to use a throttle body with a really big bore but, in fact, that is the wrong approach. The throttle body runner is, in effect, a tube supplying a plenum and can be an effective Helmholtz resonator. If the throttle body runner is of the right diameter and length, Helmholtz tuning can be used to great effect. What this means is flow derived from an efficient but smaller throttle body runner is far more effective than a bigger less-efficient runner, even if that bigger runner does flow more air.

4-1. Although these manifolds look very similar, the results they turned in varied considerably. Understanding what constitutes a good intake helps greatly to save money on an intake that delivers very little, if any, of what you are looking for. If you have the time and equipment, it’s better to cut open and modify the stock intake. Let us not lose sight of the fact that air is much heavier than often realized. When a cylinder draws air, the velocity of the air and its mass within the throttle body runner can be a significant source of kinetic energy. We can either dissipate it as heat in an inefficient runner or use it to boost pressure in the plenum. Obviously, the second option is our number-one choice. Basically, a Helmholtz plenum works like a high-velocity piston in a tube, compressing air in a closed chamber in front of it. When all of the piston’s kinetic energy has been used up compressing the air in front of it, the piston reverses back down the tube. When another induction pulse occurs, the piston (which in reality is a slug of air) is sucked back in, builds speed in the tube, and the whole process starts again. Added to this is the natural resonant frequency between the plenum and the tube.

By adjusting the three variables that dictate the frequency, we can boost the pressure in the plenum at a time when the runners are out of phase and doing the reverse of ramming the cylinders. This is usually a low-speed deal and, if such a system is tuned right, it can be very effective at pumping up low-speed torque. Although it works well with four cylinders attached to the plenum, the best is to have three cylinders so attached. All eight cylinders of a V-8 reduce the effect to minimal proportions, but if one bank of a two-planed crank V-8 operates from one plenum and the other bank from another, the Helmholtz effect here can be very pronounced.

This is the car used for the intake tests. After each change of intake manifold, the fuel and timing were checked and re-optimized as necessary.

The butterfly is a big flow impediment. A streamlining exercise netted more horsepower, with no downside than the manifolds shown nearby. To produce a system like this, the starting place is to calculate the intake diameter required. Although not yet totally conclusive, my dyno testing to date has indicated that 170 to 180 feet per second air speed in the intake tube at peak power is about as high as needed for optimal results. This leaves us with the following formula:

If you are working in metric, then use:

Where: D is Displacement in CC or inches V is Velocity in ft/sec or meters/sec VE is Volumetric Efficiency percent

Length and Volume

So far, we have arrived at a working formula for a subject known for its mathematical complexity. Unfortunately there are, at this point, some serious obstacles to achieving mathematical simplification of the plenum design yet still achieving near-optimal results. To be of the greatest benefit, a Helmholtz resonator needs to boost power just below the RPM at which the intake runner length comes into tune. Although it can be made to work with all eight cylinders attached to one plenum, it does so at about 2,000 rpm and with a much smaller intake than predicted here. For best operation, there must be a well-defined stop/start flow to the intake. This means the plenum needs to be attached to a maximum of four cylinders, but three is close to optimal. A V-8 with two induction pulses just 90 degrees apart looks (to the plenum) like a threecylinder engine with one of the cylinders bigger than the other two. So for this system to work on a V-6 or a V-8, two plenums are needed. The intake runner must then be sized for half the engine’s displacement.

By suitably sizing the plenum and intake runner, the Helmholtz effect can develop a useful degree of plenum pressure. This relies on the velocity in the intake tube just prior and during an induction stroke. Dimensional Determination In practice, a Helmholtz resonator does not act exactly as such on an engine because the engine keeps on drawing air from the system. Under these circumstances, the formula for determining what is optimal is extremely complex, so a more cut-and-shut method is called for. Let’s first deal with plenum volumes. This selection can be tricky; the intake runners need to be able to draw air out of the plenum efficiently. This often changes what can be effectively made. But as a first approximation, a plenum having a volume of about 20

percent more than the combined displacement of the cylinders being fed is a good starting point.

The first move on this 5.0 Mustang induction system was to fix the abrupt dog leg (1). This was done by first welding (2) because there was insufficient metal to port as needed. The underside of this manifold had considerable area exposed to hot oil, which put a lot of heat into the intake charge. This was fixed (3) by foam-filling the underside and making up an undertray to contain the foam and stop oil-induced deterioration. The thermostat/crossover passage is normally cast joined to the walls of the front runners. This heats them up, so the runner was separated (4). The entrance (5) into the runners on these Ford GT40 manifolds was an abrupt sharp edge. Flow was increased considerably by rounding them as seen here. Silicone sealer was used to seal the exposed foam on the top side of the intake (6). All this effort was worth 30 ft-lbs and 40 hp! Now for the length required. For an engine turning 10,000 rpm, a plenum intake runner length of about 7 to 8 inches is required. For every 1,000 rpm below that, increase the length by 1¾ to 2 inches.

Intake Runner Lengths and Diameters We now get onto the tricky subject of runner dimensions. To appreciate why these two dimensions are so important, you need to understand a few basic facts. First is that air is heavier than you may think (an average school gym contains about 40 tons of air!) A suitably high port velocity helps ram the air into the cylinder at the end of the induction stroke. As the valve closes, the air piles up, creating a positive pressure that, during the last few degrees the valve is open, helps push the last few CC of air into the cylinder. On a well-tuned system, the pressure just before and at valve closure can reach 7 psi above atmospheric pressure. Also (if the exhaust is tuned right for the RPM involved), a very strong negative-pressure wave can arrive at the exhaust valve during the overlap period. This low-pressure wave is communicated to the intake through the combustion chamber. This low-pressure wave can also be very strong. A well-tuned exhaust can pull 4 psi of vacuum here, and an optimally tuned race system, as much as 7 psi. This, on an all-out pushrod V-8, can result in the intake charge moving into the cylinder at speeds up to 80 mph before the piston has even started on its way down the bore. To make all this happen at the desired RPM, the intake needs to be a certain length. The following formula gives the required length:

Where: L = Intake Length (from the intake valve to the open end of the intake runner) ECD = Effective Cam Duration V = Pressure Wave Velocity (about 1,300 ft/sec) RV = Reflective Value (usually 2 but, for a tuned length for lower RPM, can predict an impractically long intake length. In this instance, a less-effective but more-convenient RV of 3 or even 4 can be used.) D = Diameter of Intake (at the end of the intake tract just before any entry flare)

Here is the finished engine utilizing the induction system on page 38— 402 ft-lbs and 443 hp from a totally street-drivable 302. Along with that went good fuel mileage too. ECD is an assessment of when the valve is open sufficiently for some useful activity to be occurring. For a typical engine, subtract 15 degrees from the 0.020 tappet-lift duration to arrive at a good approximation. (By the way, cam manufacturers’ catalogs have all these). As an example, let us assume we want to tune the intake length to 8,000 rpm for a fuel-injected four-cylinder engine. ECD at 0.020 = 285 degrees V = 1,100 ft/sec RPM = 8,000 RV = 2 D = 2.25 inches Inserting these numbers into our equation, we have:

This equals 15.47 inches.

This Chevrolet LS6 engine tops the 750-hp mark at 7,400 rpm and features an 11-inch runner. From this, we subtract half the inlet diameter, which is is 2.25 inches in our example. So, 15.47 – 1.125 gives us the final length of 14.34 inches.

V-8 Intake Manifolds So far, we have looked at the relatively simple Helmholtz resonator plenum-style intakes and independent runner intakes required for evenfiring inline engines of 2 to 6 cylinders. However, the most popular type of engine to modify is still the two-plane crank V-8. By having two rods on each journal and four journals, each spaced 90 degrees apart, the two-plane crank makes a V-8 function as two V-4s—not two inline 4s. This layout produces crank firing angles of 270, 180, 90, 180, and back to 270 to repeat again. It is these angular differences that give a two-plane crank V-8 its distinctive throbbing exhaust note, leaving the impression of fewer RPM than is actually the case.

This typical, stock two-plane intake for a small-block Chevy utilizes a Quadra Jet carb. These feature small primary barrels and large secondaries. Because little attention was paid to port routing, the airflow on these intakes is very poor.

Here, the more direct routing of the runners of a single-plane V-8 intake manifold can clearly be seen. This, and the ability of each cylinder to utilize all four barrels of the carb, gives this style of manifold a clear topend power advantage.

The routing of the runners for a two-plane intake can be seen here. The numbers in the ends of the runners are the typical flow numbers for a stock intake. After a week or so, porting the flow figures posted outside the runners was achieved. Although substantial gains were seen, these still fell way short of a highly developed modern two-plane.

This layout has given rise to two principal types of intake manifolds: the two-plane, and the single-plane. A two-plane intake divides the engine so that the runners joined to either half of a 4-barrel carb draw 180 degrees apart. For this reason, this type of intake is also known as a 180-degree intake. Its principle advantage is that there is no overlap of induction cycles; the interaction of one induction stroke on the next, and its negative impact on idle and cruise vacuum, is negated. The down side is the runners must follow a more tortuous route from the carb to the intake port and the effective carb flow seen by any cylinder is halved.

This RPM Performer Air Gap features runners separated from the lifter valley to cut heat transfer to runners. The two-plane Air Gap is lightyears ahead of factory two-plane intakes. Holley, Dart, and Professional Products also produce similarly effective two-plane intakes. By contrast, a single-plane intake seeks to produce the best results at WOT and at higher RPM. With this type of intake, all eight cylinders draw from a common plenum. This not only means a more direct routing of the runner, but also each cylinder in effect “sees” all four barrels of the carb. An extension of this type of intake is the tunnel ram (see page 45).

The streamline form of a modern high-tech two-plane intake’s runners can be seen here. This is a challenging packaging problem that can take up a lot of time on the flow bench and dyno. Producing an effective two-plane intake is no easy task. Although no factory V-8 has been produced with this type of intake since the mid 1980s, it continues to be popular because, conceptually, it can be very effective for a high-performance, street-driven V-8. Since the early 1990s, considerable developmental effort has led to some truly effective two-plane intakes that deserve credit for their functionality. These designs owe their success to countless hours on the flow bench and in recent years to the use of computational fluid dynamics. Because these modern intakes have so much more airflow capability than earlier designs, it has become necessary to reevaluate just how much carburetion they need for optimal results. In essence, we see that these intakes thrive on much greater carb CFM than is traditionally accepted. If we combine that with the fact that each cylinder only sees half the carb’s CFM, then it becomes clear that these high-efficiency two-plane intakes should theoretically require much more carb CFM. This, in fact, has been borne out by my own dyno testing. By adopting this “bigger than generally accepted” philosophy with high-efficiency twoplane intakes, I have been able to produce well-mannered street motors with very reasonable vacuum for brakes, idle, etc., and outstanding top-end numbers for those fast times on the drag strip. Passing the 500-hp mark with a two-plane has historically been a difficult target to reach. However, using a carb with more than 1,000-cfm capacity, I have had street-drivable Chevys of a little more than 400 inches displacement produce in excess of 550 hp.

Much can be learned using computational fluid dynamics to visualize what is happening to the air in terms of velocity and pressure as it passes through the runners of a two-plane intake. On the left is pressure. Red is ambient pressure and yellow, transitioning through green to blue, is below-ambient pressure. On the right is velocity and its direction. Blue is the slowest, transitioning through green and yellow to red (the fastest). Note the high speed (red) at the inside turn coincides as expected with the lowest pressure.

Tests of many combinations of off-the-shelf, high-performance two-plane manifolds and carburetor sizes have been run in my shop. Here, I am being assisted, as is so often the case, by crew chief/racer Mervyn Bonnet. Our tests clearly show that an efficient two-plane intake typically responds to carb CFM numbers better than its single-plane race counterpart, which runs counter to popular belief.

These tests on a 383-ci small-block Chevy with 10.5:1 CR clearly show that a larger carb CFM pays off in terms of increased top-end output with little or no penalty at lower speeds. These numbers also show the Edelbrock manifold produces far better numbers at far higher RPM than Edelbrock claim.

Here is a 406 built in my shop in about 2006. With a 10.8:1 CR as delivered by the flat-top pistons and basic-ported Dart aluminum heads, this solid-roller cam street-driver produced 558 hp at 5,900 rpm, and 549 ft-lbs at 4,300 rpm. That’s about typical of a good 482-ci big-block Chevy but for less money and about 150 pounds less weight!

This Weiand Team G single-plane intake is a common representation of this type and design configuration. As simple as the layout may look, there are many factors to take into account in order to achieve top-notch results. Among these are runner area, length, taper, plenum volume, and so on.

If hood clearance is not an issue, the runners can take a very direct approach to the cylinder ports. This can produce a very efficient intake

manifold that, as far as the cylinder head is concerned, appears to have near-zero flow restriction. Single-Plane Intakes Without a doubt single-plane intakes are the favored style for all-out performance. Using one generally improves the power curve from about 4,200 rpm on up. With the ability to have a superior routing from the plenum to the port runner, a single-plane has a far better chance at producing big top-end numbers, but success is by no means guaranteed. Factors, such as runner area, length, taper, plenum volume and their effect on mixture quality, and fuel distribution, all must be dealt with. If any significant design error exists in these areas, the effect on power can be relatively big to catastrophic.

In an effort to get lengths, areas, plenum volumes, and fuel distribution right, what may seem like extreme measures are taken in terms of plenum design. An example serves to make the point here. While doing some intakemanifold-to-cam compatibility testing for a well-known intake manifold manufacturer, I found that one of the single-plane intakes being tested was sensitive to a small change in the plenum to the tune of a 30-percent drop in output! That said, often the changes, with experience, can for the most part, be somewhat intuitive. The factor that can often be the most problematical

is fuel distribution but even that, with experience, can, even if time consuming, be second nature to establish a fix. As mixture equality between cylinders is approached, the final steps to achieve it can be made by “stagger jetting.” This procedure, without a wideband O2 unit in each exhaust, is hard to do without a keen eye for plug reading. It is a skill that, without such instrumentation, may take you years to perfect, but it is a skill worth developing if your field of endeavor is likely to be centered around single-plane, four-barrel, V-8 induction systems.

Although it presents hood-clearance problems, raising the plenum height above the cylinder heads allows for a longer and more direct runner for the inner pair of cylinders in each engine bank. By adopting this measure, compared to many of its contemporaries, this Parker Racing 5.0 Ford intake was able to produce as much as 25 hp more from a 500-hp 5-liter Mustang motor.

Although a two-plane intake is far better for torque output below about 4,000 rpm, there can be occasions when the low-speed output of a biginch engine produces too much torque to successfully put it to the ground in first gear. This 505-ci big-block just starts to fall into that category. When this is the case, a single plane can effectively be used to limit lowspeed torque while benefiting top-end power.

Staggered jetting in the carb can usually fix minor fuel distribution problems. For Edelbrock’s Super Victor intakes, when regular symmetrical jetting is close, I usually find that rearranging the jetting (as

indicated here) produces an extra 5 to 7 hp on a nominally 525-hp smallblock Chevy. Spacers Plenum volume is a significant factor toward making optimal output in the desired RPM range. Fortunately, this is easily adjusted by means of a carb spacer. These can come in a variety of forms. The simplest is the open spacer, which simply acts as a means to extend the plenum volume by raising the carb by 1 or 2 inches. The next most common is the four-hole spacer. These holes can be of a parallel-wall design or they may be of a contoured-form design, intended to improve the airflow as it exits the carb. Contoured spacers can add 10 to 15 cfm to the carb’s airflow capacity. For the record, big-block Chevys using most of the commonly available Dominator intakes respond well to the use of a spacer of as much as 2 inches, even when an already large plenum exists.

This 350 stroker (408 ci) ran well (as seen here on a Super Victor) but was good for another 6 hp when a 1-inch-profiled discharge, four-hole spacer was added.

This single-plane Dart intake equipped with a 4500-series 1,200-cfm AED Dominator responded well to the semi-contoured spacer seen here. Mounted to a 496-ci big-block, it produced an additional 12 hp.

You may have heard it said that a tunnel ram intake is a race-only item but it’s not so. As long as the budget and hood-clearance constraints don’t relegate it to the non-starter category, then the tunnel ram can be looked on as a very viable street-intake system. Examples prepped in my shop such as this thermal-managed setup (the runners are thermal-barrier

coated and foam-insulated underside) have allowed me to produced 520 ft-lbs and more than 630 hp from a 10.5:1 pump-gas 383 street-driver.

A tunnel ram intake usually indicates that maximum performance is the goal. This being the case, time should be taken to do a reasonably-wellexecuted port matching job, as a minimum. The example shown here has been fully ported, and a thermal barrier was applied to the inside. On the outside, a powder-coated finish has sufficient shine, even in red, to help reflect heat—but a better solution here is a chromate ceramic finish. The water jacket also has an insulating finish applied to its internal walls to cut heat conducted from the water to the runners. The underside is foam insulated.

Tunnel Ram Intakes A tunnel ram intake typically utilizes two 4-barrel or four 2-barrel carbs that align each carb barrel directly over an intake runner. These runners feed from a suitably sized plenum so that, at WOT, the plenum appears (to each intake runner) to be open air that has been pre-mixed with fuel. The fact that all eight barrels of carburetion, each of relatively large CFM, can be seen by any individual cylinder means the plenum runs at barely below atmospheric pressure. The absence of a butterfly within the runner also means that the runner is uncompromised both in terms of airflow and pressure-wave reflection. This makes the tunnel ram the number-one power producer when it comes to manifolds for normally aspirated engines. It is often said that a tunnel ram is a race-only setup but, in practice, this is not the case. Assuming that the budget and hood lines stretch to

requirements, a tunnel ram installation produces a low-speed output that is midway between a single 4-barrel-carbed two-plane and a single-plane intake. Assuming a stout but nonetheless streetable cam, the tunnel ram starts to outpace the two-plane intake at about 4,200 rpm, and the singleplane intake at about 5,000 rpm. After that, it really is no contest. If a motor makes 550 hp on a 4-barrel-equipped, single-plane intake, it should make at least another 25 hp on a tunnel ram and turn an additional (and useful) 300 to 400 rpm. It may not be the cheapest of intakes to build, but the results are worthwhile. Tunnel ram installations intended to use carburetors and a pair of 4barrel injector throttle bodies can be used to equally good effect. Although cast tunnel rams are common, the casting process does make the whole manifold assembly somewhat heavier than might be desired for an all-out race vehicle. This has brought about the popularly used sheetmetal intake so commonly seen on Pro Stock engines and the like. Indeed, the sophistication of specialist-built, custom tunnel ram intakes has taken the state of art in this field to levels hardly conceived of—even in Formula 1—just 15 years ago.

One of the Dominator carbs has been removed to show the internals of this Ultra Pro Machining Pro Stock–style intake. When fitted to a twovalve pushrod engine, power eclipsed the 165-hp-per-liter mark. All-out Pro Stock engines, as of 2010, are right around 175 hp per liter. This has been achieved by virtue of volumetric efficiency figures in excess of 115 percent and compression ratios up to 17:1.

Although it may not at first seem like it, this Chevrolet LS6 intake manifold is conceptually a tunnel ram with lengths intended for an RPM range from 2,000 to 5,500. The throttle body, which mounts on the front, feeds a plenum in very much the same manner as a tunnel ram. All the long, curved runners draw from this. This intake has proven to be highly functional.

CHAPTER 5

SUPERCHARGING

5-1. We have already looked at preserving air density. Here, by one mechanical means or another, we focus entirely on maximizing air density arriving at the cylinder. Let’s focus on the word “supercharge” for a moment. In the context of this book, it means to fill the cylinder above and beyond the displacement that it ordinarily has. For instance, if a cylinder has a 50-ci displacement and, by one means or another we cram 75 ci into it, then we can say it’s not just “charged” but “supercharged.” Chart 5-1, shows that supercharging falls into the “Maximize Air Density” category. The so-called normal situation is to not lose any air density that exists naturally. This is the non-supercharged situation and is

commonly called naturally aspirated, or NA for short (the F1 guys often called them Atmo Motors). A supercharger is a device that has the ability to overfill the engine’s cylinders. But why use a supercharger? Why not just design the engine with greater displacement in the first place? That’s a good question, with a lessthan-obvious answer. The answer is very much a question of component size and geometry. An analysis shows that a piston/rod/crank mechanism is a very effective means in all respects at taking pressure energy and converting it efficiently into rotating motion. It is not only mechanically efficient but also relatively effective in terms of size and component weight necessary to do so. What a piston/rod/crank assembly is not so good at is moving air at an atmospheric pressure of typically 14.5 psi absolute (pressure above a total vacuum) from outside of the engine into the cylinder. For this job a piston/rod/crank assembly is a heavy and cumbersome mechanism in terms of component size for the mass of air induced. For a supercharger to justify its existence, it has to be (size for size) far more effective at moving significantly greater quantities of air from point A to point B than a piston/rod/crank mechanism. As it happens, there are plenty of ways to do just that. A good example is to compare the size of a turbo compared to the engine it’s feeding. What we find is that if a turbo is about one-twentieth of the size of the engine it is on, it can pass as much as four to five times the volume of air.

Cup Car engine builder Billy Fisher put together this 582-ci big-block Chevy. It features a traditional BDS Roots-style blower system paired with fuel injection. Power potential, even in streetable form, with good heads and valvetrain, is 1,250 hp along with a high intimidation factor. What this all means is we can let the crank/rod/piston assembly do the job it does so well—converting pressure energy from the combustion process and turning it into usable power—while the supercharger takes over the function of inhaling air more effectively. By adding a supercharger to an engine, we are making it act as if it had far more cubes. This is because the supercharger, not the basic displacement of the long-block assembly itself, determines the magnitude of the induction process.

Supercharger Types Superchargers fall into two fundamental categories: the positive displacement type, and all the rest. Almost certainly the best-known positive displacement supercharger is the Roots type. This unit, developed in England during the mid 1800s by the Roots brothers, was intended to pump air into deep mine shafts. However, during World War I, it saw use on aircraft engines as a means to boost power at altitude. Success here subsequently led to its use between the wars on Mercedes and Auto Union F1 cars as well as the highly successful British ERA F2 cars. The Roots blower really came to the fore when the likes of Don Garlits, and a few other leading Top Fuel drag racers of the late 1950s and early 1960s, figured they could go faster with a big Roots blower atop the motor —and they did. Back then, because there was not much else to choose from, the blower of choice was the GMC supercharger used on GM’s two-cycle diesels. The sizes most commonly used were the 6-71 and the 8-71. These big blowers could puff about 20 to 25 psi or so into a Chrysler Hemi and bump the output to what was then an incredible 2,500 or so hp and about 3,200 ft-lbs of torque. Today, we see many superchargers that owe their heritage to the GMC range of blowers. A few examples are those from Holley/Weiand, Edelbrock, Magnuson, BDS, and Eaton. When taking a casual look at this type of supercharger, it is easy to assume it draws the air into the middle of the rotors and passes it down into the manifold. In reality, this does not happen. The best way to see how it works is to look at a small-block Chevy oil pump, because this is, in fact, a mini Roots-type pump. From its inception to possibly as late as the 1980s, the big drawback with the Roots supercharger was its relatively inefficient (about 55 percent) pumping characteristics. In this quarter, real credit can be given to Jerry Magnuson, one of the world’s leading supercharger designers. His extensive work over a number of decades has brought about dramatic improvements to the Roots-style positive displacement blower. This has been achieved over the entire operating range and duty cycle, to the extent that it is closely comparable to a turbine- or centrifugal-style supercharger. To put that into prospective, he has, in the last 30 years, done what engineers in the previous 120 years largely failed to do. Today, 98 percent or more of superchargers use

Magnuson technology. In other words, your modern Roots-style supercharger is not the one your father knew.

The simplest Roots-style super-chargers are still much like the Chevy oil pump. However, the high-efficiency Magnuson/Eaton units (Magnuson shown here), have evolved dramatically. Instead of being straight, or nearly so, the rotors have a steep helix and are now classified as Hybrid Roots/Screw compressors.

A good example of a Roots-style pump is seen here in the form of a Chevy oil pump. The medium being moved enters at the top and is carried around the outside of the case between the gear teeth. As the gear teeth re-mesh on the lower side, the medium being pumped is squeezed out from between the gears and discharged out the lower side of the case. The mechanism is very basic but, in this form, lacks efficiency.

Here is a 2010 high-efficiency design from Magnuson. Instead of the classic top-surface air entry, the air enters onto the rotors at the front of the unit. In addition, a vacuum-referenced bypass butterfly (arrow) equalizes the supercharger inlet pressure with manifold pressure at partthrottle cruise. This virtually eliminates all parasitic losses other than the unit’s bearing losses. This means the impact on around-town and cruise mileage is virtually zero.

The form of the case discharge into the manifold has also been refined greatly. The discharge port shape (arrow) is no longer a straight slot, but a shape to compliment the rotor compression characteristics.

Mounting a modern high-efficiency unit, such as this Magnacharger, on an intercooled base/intake manifold, leads to higher power on otherwise lower-octane fuel.

Drive for the inlet turbine can be either by a belt-driven pulley (shown here) or by an exhaust-driven turbine, such as used by a turbo charger. Although the most common the Roots-type supercharger is far from the only positive displacement type available. Among the many others in this category are the Zoller vane-type supercharger, the novel (and supposedly very efficient) VW G-Lader, and the Whipple-style screw compressor. Of these, only the Whipple unit, also a high-efficiency unit, is in any kind of aftermarket volume production.

Seen here is a turbine-style supercharger (this unit is from ProCharger). The turbine seen through the inlet can typically spin up to 60,000 rpm.

This ProCharger unit is used with a blow-through carb-induction system. The ProCharger blower shown here is big enough to supply the air demand for an engine of 1,600-plus hp.

Turbos and Centrifugal Turbine Superchargers Barring leakage, a positive displacement supercharger moves a certain defined amount of air per revolution. A turbine, or centrifugal supercharger (as they are more usually called), develops boost by imparting motion into the air. When the rapidly moving air meets resistance, the slowing of the air converts the air’s kinetic energy into pressure energy. For that reason, the boost and airflow throughput of a turbine supercharger is closely linked to the characteristics of the engine it is feeding.

This graph shows how a nominal 10-psi boost-curve requirement at high RPM can differ radically at low RPM depending on the type of supercharger used. Although the reasons for choosing one type over another are many, the basic difference comes down to the shape of the power curve required. Essentially, a positive displacement supercharger produces more low-speed boost, while a turbine-type supercharger can develop more high-speed boost. Here is an important point concerning boost: It is all too easy to suppose that a supercharger’s whole existence is to develop boost but, as obvious as that may seem to be, it’s not quite true. If the goal was the biggest boost numbers possible, then welding the intake valves closed would do just that, but then the engine would not produce much power. The real job of a

supercharger is to move the greatest mass of air through an engine, which then must subsequently use it effectively. A turbine supercharger can move a great deal of air very effectively. Also, since centrifugal superchargers’ mode of operation does not involve so much beating around of the air, it can do so with high efficiency when optimally designed. But there is a downside; a turbine-style supercharger is speed sensitive. As the turbine speed increases, the boost possible also goes up with the square of the RPM. If you double the turbine RPM you get, in round terms, four times the boost. As far as the engine’s torque curve is concerned, at low speed there’s minimal boost, which means low torque. Assuming a supercharger is sized right for the amount of boost required for the midrange, we find that the boost goes out of sight at the top end. Although the trend is somewhat set in concrete, there are partial remedies to this lowspeed (low boost) situation. In short, a lot of work in the past 50 years has brought about a number of improvements that minimize this characteristic.

Boost Curve Shapes We can best see how the “too little low-down and too much high-up” situation has been addressed by looking at a typical turbocharger installation. In essence we have an intake turbine driven by a second turbine in a housing through which the engine’s exhaust is fed. Simply put, the exhaust energy is driving the intake compressor turbine. The general principle here is to size the intake turbine so that it is a little oversize—without any limiting device, it produces more high-speed boost than needed. This over-sizing means boost to the engine starts sooner but, if left to its own devices, provides too much at high speed. To prevent this, the exhaust side is equipped with an exhaust wastegate. When the boost reaches a certain predetermined level, the wastegate opens and bypasses any exhaust excess for the job of spinning up the intake turbine. This, along with impellor design characteristics and inlet sizing, helps spread a turbo’s boost range down into the lower-RPM band.

A Weiand blower on a mild 350 small-block Chevy added about 110 ft-lbs and 110 hp over that expected in NA form. The output produced is similar to a mildly modified 454 big-block, but without the big weight increase. For a mechanically driven turbine supercharger (ProCharger, Vortec, Paxton, etc.) of the type originally pioneered by Paxton 60 years ago, we find the situation a little different. Because the drive ratio between the supercharger and the engine is fixed (Paxton actually developed a variabledrive unit used on Ford Thunderbirds in the late 1950s), limiting top-end boost while trying to enhance low-speed boost comes down mostly to impellor design and overall unit sizing. In that area, significant strides have been made since the mid 1990s.

Supercharger Selection

At your local library, in the reference section, you can find books on performance going back as far as the 1930s. A look through these might just give you the impression that the heyday of superchargers was in the 1930s through to maybe the mid 1950s. The reality is the heyday is from now into the foreseeable future.

This Magnuson kit for a 2000-model-year 4.8-liter GMC Sierra truck was installed and produced very strong results. Since then, Magnuson has made further upgrades in blower efficiency.

Here, my 2000 Sierra is being dyno tested on Custom Performance’s (Charlotte, North Carolina) Dynojet dyno. Final results netted more than 380 RWHP. If you are concerned about fuel economy, and the possibilities it has to negatively impact building big power, let me put your mind to rest. From

the auto manufacturers’ viewpoint, there are numerous avenues toward employing a supercharger to enhance both performance and mileage. That alone more or less guarantees their continued use for a long time to come. If your intent is to get additional power in the form of a bolt-on supercharger system, then your only real problem is making a decision as to which of many excellent systems you should use. If It’s Low Speed You Want As of 2010, I own a 4.8 GMC Sierra truck and, like many truck-owning performance enthusiasts, I want it to perform like a sports car for day-today driving. But I also want it to carry out its work functions—towing a loaded trailer and such—better than stock. When I am towing the racer to the track, that 4.8-liter V-8 can be hauling a payload of as much as 10,000 pounds. To do this effectively, the engine needs real low-speed torque along with a good fuel economy capability. About the time I was ready to start modifying this truck’s engine, I had a fair amount of experience with Holley/Weiand, Magnuson, Edelbrock, and Whipple kits, but mostly with the Holley/Weiand and Magnuson stuff. My simple blower build done in the mid 1990s produced, on a mild-cammed 350 with pocket-ported aluminum heads, had some 541 hp and 545 ft-lbs output using a Holley/Weiand blower. This was a pretty good showing for what is essentially one of the lower-cost installations on the market. As with most trucks used for towing race cars, the distances involved can be quite substantial, so fuel mileage was important. This factor alone caused me to re-visit Magnuson supercharger installations available at the time. Since my last go-around with a Magnuson supercharger, the design had been modified with a cruise-mode bypass valve. This valve serves to reduce parasitic losses to barely more than that taken to spin a couple of sets of roller bearings. So with serious heavy-duty towing and mileage in mind, I selected a bypass-valve-equipped, positive displacement, intercooled Magnuson kit. Because the Magnuson kit had charge intercooling via a water-to-air intercooler between blower and block, it could manage 8 psi on a 9:1 engine without being detonation prone or octane sensitive. To make the most of this supercharger package, the stock 4.8 heads were ported and a Gale Banks exhaust system was installed.

Results were very much what I expected. The quarter-mile performance turned in a result just short, by a truck’s length, of that produced by a nearstock (K&N cold-air kit only) 2004 five-speed Mustang GT. Freeway mileage was just shy of the 21 mark while a consistent 17-plus was seen about town. That was almost unchanged from stock. When towing sensibly, mileage was surprisingly good. Most of my tows involve 90 percent or more freeway travel. If speed was kept to 60 to 65 mph or so, and with a 6,000-pound trailer load, I saw low 14s for MPG. The results of these tests strongly suggest that for a good balance between performance and mileage, a smaller engine with an efficient supercharger is the way to go. In 2008, Magnuson introduced yet another significant step forward in rotor design. This newer design in photos on page 48 is essentially a hybrid Roots/screw compressor. The efficiency figures rival those of a turbine-style supercharger but have the advantage of boost right off idle. This in itself allows an auto manufacturer to use a smaller engine because the low-speed boost totally compensates for any torque reduction due to the smaller displacement. At this point we can say that (in the main) off-idle and low-speed torque, be it for a truck or a true street car, means looking first at a positive displacement blower. If low-speed torque below about the 2,000- to 2,500rpm mark is not an issue, then a whole slew of supercharger types become a viable bolt-on option. That is what we next consider.

The baseline curves (black) are for the truck, stock other than a Gale Banks exhaust system, which added some 12 ft-lbs and 14 hp. The ported heads increased output as shown by the blue curves. Adding the Magnuson blower setup with 7- to 8-psi boost to this combination increased output by 120 hp and 123 ft-lbs of torque. Centrifugal Superchargers In most instances, the simplest installation is a belt-driven centrifugal supercharger rather than a turbo. Although they may lack the low-speed capability of a positive displacement unit, they can usually (but not always) deliver a cooler boosted charge, thus allowing more power to be developed before an intercooler or water injection becomes necessary to eliminate detonation. Regardless of the cooler charge advantage, most installations still cool the boost with an intercooler. When combined with intercooling, the results can be little short of spectacular. Here’s an example to illustrate the substantial difference that can be made: Bolting a 10-psi boost ProCharger unit on an otherwise-stock 4.6-liter, three-valve 2007 Mustang produced a rear-wheel-hp increase from 260 to just over 460! Here, we have an engine boosted to 68 percent above atmospheric pressure but making 75 percent more power.

At first glance, a Whipple supercharger bears much resemblance to the classic Roots-style supercharger. A check of the interior workings might also support this; the rotors resemble those of a Roots supercharger but with a very high helix angle. The reality is that it belongs to a class of compressor known as a screw compressor. This has a male and female rotor, and the way they intermesh produces a far more efficient means of moving and compressing the air that is drawn in.

First, imagine these two rotors housed in a case with an opening at the top back and a discharge at the lower front. Starting from the left, imagine that the groove on the right-hand rotor is full of air. As the lobe of the left-hand rotor rotates into this groove (middle), it can be seen to close off any exit to the back where the intake is. This traps air in the groove and, as further rotation takes place, the air trapped in the groove gets pushed forward by the left pair of rotors. This squeezing action, moving and compressing the air, takes place very efficiently with figures rivaling those of a turbo. How does this work out? There are two factors at work here: keeping the charge temperature down by means of an efficient supercharger, and intercooling it. In addition, we now have a situation where the boosted intake charge tends to blow much of the residual exhaust in the combustion chamber out through the exhaust valve. This leaves us with an intake charge equal (more or less) to the cylinder displacement, plus that of the combustion chamber. When it comes to big mid- and top-end numbers, a centrifugal supercharger is the way to go. Now it becomes a question of opting for a mechanical drive (belt driven) or an exhaust-driven turbo setup. Just as in making a decision between positive displacement and centrifugal superchargers, there are pros and cons for each type of drive. If the supercharger is mechanically driven, there is no boost lag as the throttle

is opened. This implies that a turbo will have boost lag, but that is not necessarily so (which I discuss on page 54). In addition to possible lag, there are other issues that need to be considered. The first of these is that, unlike turbo-driven systems, a mechanically driven supercharger produces much less engine compartment heat. That said, a turbo setup, which recovers some of the normally lost exhaust energy, still ends up to be the top contender for total output. Another advantage of the turbo is that it can be sized to come on sooner, with the top-end boost limited by the wastegate. At the end of the day, both types produce dazzling performance from most otherwise-stock engines. The performance is so sufficient that the problem becomes one of getting all the power to the ground. When purpose-built from the start with the intent to supercharge, the power potential is little short of staggering. As of 2010, I have helped on the build of turbo engines making about 10 hp per cube (just over 600 per liter), and that is without a budget of any real consequence. With a much bigger budget to work with, I have seen some factory development projects that have pushed the turbo envelope on a fourvalve engine to 13.5 hp per cube (835 hp per liter).

The Squires Turbo System’s rear-mounted installation is definitely less than conventional. A similar remote installation was used for the Republic P-47 Thunderbolt, a very successful fighter airplane of World

War II. When space is limited up front, this makes for a viable option. Another benefit is that underhood temperatures are also reduced.

Here is an example of a twin-turbo system from Air Power Systems in Australia. As indicated by the arrow, the turbos are situated low and on either side of the engine. This install requires a tubular subframe to make everything fit. Although an extra cost is involved with the subframe, cutting some 70 pounds from the front helps reduce the inevitable traction problems that big power numbers bring about. But let’s drop down the budget ladder a little here. A couple of moreaffordable projects I have worked on put a realistic perspective on what can be achieved. A ProCharger-equipped, 351 Windsor-powered Mustang turned up the dyno rollers to 850 hp and at that point just smoked the tires. I estimate it was still more than 1,000 rpm from a peak power number that could well have been in the four-figure range. With a turbocharged smallblock Chevy with a Performance Techniques installation—same deal. Both of these vehicles most certainly exceeded 1,000 rear-wheel hp. Even on the grippiest road tires available, neither of these cars hooked up at less than 120 mph! That’s on a dry road; on a wet road, you would need all of Michael Schumacher’s driving skills just to survive.

If we are dealing with a well-engineered supercharger installation, cost, ease of installation, and personal preferences need to influence the system you choose. But let’s get our feet back on the ground here. A more bolt-on turbo kit for an otherwise-stock engine, such as an LS-series GM engine or Ford’s three-valve 4.6/5.4 unit, responds to the tune of a 200- to 300-hp increase with a boost of 8 to 10 psi. A good example here is the kit offered by Turbonetics for the 2005–2008 Mustang. This boosts rear-wheel hp from the stock 250/260 to about 500 and all done with less than 10-psi boost. Here it is worth addressing the subject of so-called turbo lag. About 1977, I was helping my friend Jim Flynn with his turbo 2-liter Pinto. It was a stick-shift setup, so the engine could not be loaded up against the torque converter prior to a launch. This meant I had to come up with an alternative way to create boost before the car was launched from the line. The idea I came up with (one that at least two other people have subsequently claimed to be their own) is very simple. If, at the launch RPM, the engine’s ignition timing is drastically retarded or even has a random spark cut, the RPM will not climb; but the throughput of air and fuel, and consequently the production of exhaust, continues. This keeps the exhaust turbine spinning at high RPM, producing boost even when the car is in neutral. The technique at the start line is to set the launch RPM, and then floor the throttle against this limiting RPM. At this point, the boost comes up while the car is still staged and ready to go. When the clutch is released, the ignition timing comes right back to its proper setting, and suddenly you have more power than you know what to do with.

This 8-psi boot Air Power Systems twin-turbo kit was installed and tested at PCM-For-Less in Mooresville, North Carolina. This system pushed the rear-wheel horsepower from a typical stock 280 to 300 up to 552. This system worked very well to the extent that we had to actually tame it down a little. I wrote about it in detail in a book I did many years ago on 2-liter Pinto engines. Subsequently, virtually all of the turbo Formula 1 cars of the 1970s and 1980s used this system. These cars were virtually fly-bywire systems, in which the throttle blades were not actually connected to the throttle pedal. Instead, the throttle pedal was a torque-demand pedal. When the driver lifted off the pedal, the throttle blades only partially closed; but at the same time, the ignition retarded by a large amount. The charge throughput did not make for any flywheel horsepower but it did keep the turbo spooled up.

Superchargers and Built Motors So far, we have mainly looked at supercharging as a bolt-on for an otherwise-near-stock powerplant. In such a situation, the kit manufacturer pretty much takes care of issues that may arise by limiting the power increase supplied, limiting boost, and including an intercooler. Now let’s move on to consider what issues you may need to address if you are looking

for competition-killing numbers for opposition annihilation at the track. There are four primary issues we need to address: detonation avoidance, control of excess heat, maximizing mass air through the system, and the effective and reliable conversion of the mass air throughput into torque and horsepower. Detonation Avoidance Combating the onset of detonation is the most important factor to deal with. The bottom line here is: If the engine is strong enough to withstand all other loads and stresses, then detonation ultimately limits power production. First, higher boost means lower compression ratios (CR), but there is a balance to be struck here. For good part-throttle fuel economy, the balance needs to be more in favor of the CR. If ultimate power is sought, the balance is toward maximizing boost. The engine’s combustion chamber, exhaust valve temperatures, and the fuel’s octane rating determine just how high the CR can go under such circumstances. Charge cooling by means of an intercooler and higher fuel octane are the two routes to success here. Control of Excess Heat As a means to combat excess temperatures and/or inadequate fuel octane, water injection has no equal. First, the octane rating of water is virtually infinite and, while boosting the octane, it won’t even come close to putting out the fire. Spraying even a huge amount of water as a fine mist into a burning air/gasoline mix serves only to reduce peak temperatures and make a lot of steam! The cooling action is effective to the extent that melting anything is totally countered. Here’s an example to give you an idea of just how effective water injection is: During a class project, we successfully ran a 17:1-compression tractor engine on kerosene (less than 50 octane) and water injection. More in line with what we are doing here, I helped build a 1,100-hp, 350 small-block Chevy that ran 35 psi and 87octane fuel. Water injection absolutely works. And if you are in the market for a system, check out Snow Performance.

Snow Performance is the biggest name by far in the water-injection business in the U.S. The system allows the Magnuson-blown Sierra (shown earlier) to run 87-octane fuel without fear of detonation or the sacrifice of any power! Controlling excess heat with water injection is a good start, but when very high output is called for, this needs to be complemented with a big radiator and piston-oil squirters. The big radiator is self explanatory but to those who may be new to supercharging’s little nuances, an explanation of squirters is in order. Here, engine oil is squirted through jets onto the underside of the pistons, thereby oil cooling them. I have used this as a procedure on all my serious nitrous and supercharged engines since the late 1980s. Many of the newer, small four-valve factory turbo engines are now adopting this method of increasing piston life. Maximizing Mass Air There are many misconceptions associated with this subject. The most notable concerns the cylinder heads. Do not fall for the myth that heads don’t need to be good because the supercharger forces the mixture into the cylinder. Ask yourself where the power to generate the extra force necessary to push the charge into the cylinder came from. Just as normally aspirated engines respond to better flowing heads, so do supercharged engines—to an even greater extent. Also note that for higher boost figures it is better to trade off some intake valve diameter for a larger exhaust diameter. At this time, I am building a 20-psi-boosted, 331 small-block Ford. The valve combination being used is 1.94/1.7 instead of the more

normal 2.02/1.6 combo. Be aware that the lower the CR used, the greater the bias toward exhaust valve size needs to be.

Superchargers and high-flow heads are a match made in heaven. Each complements the other, and the sum total of the two is greater than each one independently.

Getting the Cam Right Let us set aside the cam requirements for a turbo motor for the moment. We can then set ground rules that apply to all supercharged engines where the exhaust flow is uninhibited. The most important factor here is the lobe centerline angle (LCA). This is very dependent on the boost/cubes combination in relation to the circumferential length of the intake valve and the intake/exhaust flow ratio. When boost is applied progressively we find the optimal LCA needs to get progressively wider. Also, as is so often the case, the exhaust valve is too small for a supercharged application. It needs to be opened earlier, thus widening the LCA itself. Although not so good for the bottom-end output, the early exhaust opening can help the top end appreciably. What we are trying to avoid here is the boosted intake charge passing right through the combustion chamber during the valve overlap period and on out through the open exhaust. Another factor that needs to be taken into account for a street setup is that the cam for the job can be shorter. This effectively boosts the low-speed

output from an engine equipped with a centrifugal-type supercharger. In this situation, if the heads are good, then the supercharger takes care of the top end.

Turbo Cams If you understand what I am about to explain, you probably have a 90percent chance of saving yourself from wasting money. What you need to understand is that a cam company cannot sell you an appropriate cam for a turbo motor until they know the pressure differential across the motor. When a turbo goes into boost on the intake side, it is driven there by pressure built up between the engine and the turbo. Other than in a pulsedriven turbo system, it is this pressure that drives the exhaust turbine. Most turbo kits have an install design priority. This means they may give up power-enhancing moves, which allows for a simpler and cheaper installation. As a result of necessary compromises (or just a not-so-good design) a typical turbo kit can have an intake-to-exhaust pressure ratio of about 2:1, but that can vary from 1:1 for a really well-designed installation right through to a mediocre 3:1.

Roots superchargers respond really well to high-flowing heads. With a streetable cam, this 572-ci blown and injected BDS setup can push past

1,200 ft-lbs and 1,100 hp. Let’s assume that your turbo setup is going to be about 2:1. If that is so, then, when there is 15-psi boost in the intake, there is 30 psi of pressure in the exhaust manifold. Let’s say a cam with a typical street overlap is installed and the intake valve opens. Then, rather than the mixture entering the cylinder, exhaust passes out through the intake valve into the intake manifold. This heats the charge considerably and often increases the likelihood of detonation. Working with one of the country’s leading turbo shops, we investigated cams with negative overlap where a high-negative-pressure ratio existed. For about a 2:1 pressure ratio, we calculated that about minus 25 degrees of overlap should work, and it did—big torque right off idle, big power, and a glass-smooth 600-rpm idle. Also the engine was far less sensitive to fuel octane. When building my own turbo motors, I work hard to target a 1:1 pressure ratio across the engine. If this is achieved, the cam selection is a whole lot easier. If this 1:1 ratio is achieved (as was the case with our turbo Ford Cosworth Sierras, which won the road race championship in England), any cam that produces the best events for a normally aspirated engine also works best for the turbo setup. There is one more valvetrain factor to take care of concerning the valvesprings. Because boost pressure tries to open the valve, a stronger spring is required on the intake of a regular supercharged engine, and on the intake and exhaust of a turbo motor.

Fueling System Now that we have dealt with pushing air through the engine, it is time to consider just how the appropriate amount of fuel is added. Basically we have three options: a suck-through carb, a blow-through carb, or fuel injection. This choice dictates much when planning a supercharged installation. If we go back to just the 1980s, the fuel injection option was probably little or no better than the carbureted option. But in this day and age of precision electronic fuel injection, and budget permitting, fuel injection is the way to go with any blower setup where total output is the number-one priority. However, don’t infer from that statement that a

carbureted supercharged setup is a distant second best, because it most certainly is not.

If you read nothing else in this chapter, read what I have to say about cams for turbo applications; it could save you from wasting of money. The advantage of spending $3,000 on an injection setup is that there is zero fiddling and zero doubt about the tune when you are done. Getting the fuel curve right-on is just a matter of reading the oxygen sensor output and adjusting accordingly—that’s if the injection system itself did not actually do it for you. Going the carbureted route can save between $1,000 and $2,000, but you almost certainly need to do at least a little fine tuning. If the blower is a Roots-style unit, then the most obvious and convenient place to put carburetion is on top of the blower itself. This is the classic draw-through mode of operation. Apart from referencing the power valve to the intake manifold boost, the carb pretty much operates as normal. But with some installations, building a draw-through system is not quite as easy; centrifugal superchargers sort of fall into this category. Here, it is more convenient to have a blow-through system. Unless you have a carb built by a company knowledgeable in this mode of operation, you could encounter many problems. On a scale of 0 to 10, your carb calibration problems could easily be 5 or more—but you need zero problems. Although I am sure there are others out there, I have successfully used Holley-style, blow-through carbs, from Holley, the Carb Shop, and AED. As for what you may give

away in terms of power, you should realize the potentially large number of ft-lbs of torque that the engine inevitably makes, if it’s done right.

When boost pressure exceeds about 8 psi, it is necessary to re-evaluate the spring’s delivered seat force. A typical small-block V-8 has 3 square inches of intake valve, so even a low boost level of 10 psi backs off the closing force by 30 pounds. This has to be countered or the valve motion fails to track the lifter motion, even at low RPM.

This blow-through carb setup, in conjunction with a Vortec blower, allowed a moderately modified big-block Chevy to achieve a low-fourfigure output. The cost of going the carb route here is significantly less than fuel injection.

A blow-through carb from The Carb Shop in Ontario, California, seems to be a very popular move. Recommended by many supercharger companies, this carb, in single or dual form, can help deliver as much as 1,800 hp.

CHAPTER 6

CYLINDER HEADS We have finally arrived at the so-called cork in the system, as far as power production is concerned. As Chart 6-1 shows, optimizing airflow (and that means getting as much as possible) is the number-one goal. But as far as heads are concerned, it is not the only factor in making big power numbers. To understand the technicalities involved, we start with the need for good airflow. Without this, it matters little what other factors are imperfect; the engine makes less power, no matter what.

Optimizing Cylinder Head Airflow With the classic poppet-valve-style engine, air cannot make a straight shot into the cylinder. There has to be a bend in the port to accommodate the valve stem and, worst of all, the air must make its way around the head of a valve. It is the head of the valve and the seat on which it rests during the closed position that is the worst impediment to flow. Although I have used it many times before, Illustration 6-3 serves well to put into perspective the relative flow restriction each section of a typical production port has.

6-1. In addition to optimizing airflow, we must also address the three factors listed below the arrow, and that is a tall order by any standard. Within either port, the valve is the only component that, as it opens and closes, delivers a variable geometry. At low lift, the flow is entirely dependent on the size of the gap between the valve seat, the seat in the head, and the efficiency that it flows at. At low lifts, the geometry of the seat and the area closely surrounding it play a huge role in dictating its flow efficiency. At high lift, the port size and shape dictate the amount of flow delivered. This being the case it’s helpful to have an idea of how the flow priority changes from seat form to port form as the valve goes through its lift cycle. With the aid of a flow bench and some calculations, this can be determined; the result (shown in Chart 6-4) is that the valve seat influences flow to a much higher lift than you may intuitively suppose.

6-2. High airflow within cylinder heads is all about getting air around corners. Assuming a strong bottom end, it’s ultimately the ports, valves, and seats that dictate the power produced. In Chart 6-4 two lines cross (in the circle) and, at this point, the form of the seat is equally as important as the size and shape of the port. Although these results were from a production pushrod cylinder head, they can be quantified to apply reasonably well to any style of head—two- or four-valve; stock or race. This can be done by looking at the lift-to-valve diameter ratio. In other words, we are reducing things to proportions rather than absolute dimensions. A key proportion is a lift value equal to one quarter of the valve’s diameter. This is commonly known as 0.25D; Illustration 6-5 shows the significance. At 0.25D, the “curtain area” is exactly equal to the valve head area. By applying the same criteria, we can say that the valve seat influence, though continually diminishing, is the major influence on flow, up to a valve lift equal to about 0.18 to 0.20D. In other words, the seat is the number-one priority until the valve has reached a lift figure between 18 and 20 percent of the valve’s diameter.

6-3. Studying this illustration of a small-block Chevy intake port for a few moments shows that the most tortuous part of the air’s passage from the manifold face into the cylinder is as it passes around the valve. For this reason the form, for 1/2 inch up- and downstream of the intake or exhaust valve seat, is the most important element for producing optimal airflow.

6-4. Comparing seat and port velocities indicates the relative flow priorities of each. This cross-over point usually ocurrs at about 0.180.

At this point, you may well ask just how much difference a valve seat shape can make to the efficiency delivered at the lower lift. The answer is: a lot. Up to the 0.20D lift point, a typical production three-angle valve seat usually delivers efficiency figures starting at about 65 percent and quickly dropping to around 55 percent. Spend time (a lot of it) on a flow bench, and those figures can get as high as 85 percent and sometimes more. A point worth raising here is that there is a commonly held belief, among many successful engine builders who specialize in high-output two-valve V8 engines, that too much low-lift flow hurts power. I won’t go into a lot of detail here other than to put forward a simple explanation. If you want a highly detailed analysis as to why this is not so, then go to my porting school series at www.motortecmagazine.com.

6-5. To better understand how flow efficiency can vary as valve lift progresses, it is helpful to look at lift values in relation to valve diameter. Here we see one of the most important lift points, the 0.25D. Depending on the style of head, a number of important things happen here that explain much about the valve lift needed for maximum power.

The simple explanation is: First, the cylinder does not know how far an intake valve is open. All it knows is how much flow is being presented. Therefore, if too much flow is being presented at any given moment, it is because the cam valve events are not what the head/engine combination wants. That means the wrong cam is being used. Claiming a cylinder head has too much low-lift flow is a little like claiming your race team has too much money!

The performance thermal-barrier coatings on the Ultra Pro Machining CNC LS6 head allow us to see the seats more clearly. Seat form is as important as the ports for any performance-head development program. In Illustration 6-6 “A” is a knife edge seat (the port in the head is the same size as the seat). This gives the maximum area for air to enter the cylinder at high lift, but the sharp corner is very un-streamlined. That means even if we could seal the valve in this instance, it still would not flow well. The obvious answer to the sharp corner is to chamfer it with a 45-degree cut as in “B.” Even though the area directly under the valve is smaller, the combination flows better. If we work on the principle that some streamlining is good, then maybe more is better. However, we find that a second angle at 60 degrees beneath the 45, as in “C,” improves this even further. Finally, a small exit cut on top of the valve seat at 30 degrees further helps the situation, as shown in “D.” All this tells us that the seat and the port in the close vicinity are all about size and streamlining.

6-6. Here, we see a progression of seat design from knife-edge (A) to threeangle (D). This brief description of the developed shape of the seat does tend to make things look easy, but thinking so is a mistake. As we shall learn later, flow around a valve is not even because of the direction of its approach. This means that the simple valve seat form we have just discussed is probably okay but hardly optimal. To get a near-optimal form with any given port requires a flow bench and usually a fair amount of testing.

Valve Shrouding You may have heard of valve shrouding; it and its effects are closely tied with the seat and port design that works best. For that reason, let’s look at it now rather than later. So what is valve shrouding? It is easier to see from a drawing than to explain, so take a look at Illustration 6-7. This is an intake valve on the intake side of a combustion chamber. The point to note here is that the walls of the chamber are, for much of the valve’s circumference, close to the edge of the valve. This shrouds that part of the valve, limiting flow around that

particular section. The green line represents the radius of the stock chamber wall, and the airflow produced by this is shown by the green line on the chart in 6-8. We obviously cannot cut the chamber away where it is adjacent to the bore diameter (gray line), but elsewhere we can cut it to alleviate, as far as possible, the shrouding caused. We can look at the chamber wall as a continuation of the valve seat. At an angle of 36 degrees from the valve seat on up, the area around the valve is always as much as the curtain area. This represents a geometrical unshrouded valve.

Practical De-shrouding Illustration 6-9 shows a typical two-valve-per-cylinder, wedge-shaped combustion chamber. It is already limited on breathing capability simply because it is only a two-valve-per-cylinder design. It makes no sense to further limit this type of head’s power-production capability with needless shrouding. All this talk of shrouding raises the question as to whether it is possible to have zero shrouding within a head that still utilizes the largest valves possible. A hemi-style combustion chamber can provide just that. The reason it does so is because the valves are always moving away from the cylinder wall as they open.

6-7. On the left is a totally shrouded valve. This doesn’t happen in practice, but it does serve to demonstrate what shrouding is. As can be seen, no air passes by the head of the valve because it is totally enclosed by the chamber walls. By pulling the chamber wall away from the valve’s circumference, as on the right, valve air (green arrow) can pass freely around the valve head.

6-8. Progressively cutting a heavily shrouded combustion chamber away (as per the colored lines in the illustration on the left), showed the corresponding flow test results (shown by the graph on the right). Note that this follows a law of diminishing returns.

We can say a valve is geometrically fully un-shrouded if there is a clearance around the valve head equal to 0.20D at 0.25D lift. This equates to an angle, for the combustion chamber wall, of 26 degrees from the vertical.

6-9. We can do little to remove shrouding caused by the bore diameter (red), but chamber shrouding (green) can be minimized to good effect.

A hemi-style cylinder head has zero valve shrouding. This comes about because, as the valve opens, the distance from the edge of the valve to the cylinder wall increases from A to B.

The latest Chrysler Hemi engine is an up-to-the-minute, high-tech design. In stock form, the heads flow as well as many good aftermarket heads for the small-blocks from GM and Ford. In ported form, they rival a set of $10,000 Cup Car heads.

This small-block Chevy Dart CNC-ported head is a good example of shrouding reduced to a minimum. The red and the blue lines represent zero shrouding. Only the cylinder walls cause any shrouding.

A lot of effort went into minimizing shrouding on this D3 Ford Cup Car head from Ultra Pro Machining. The result of this and other intensive airflow work is 850 hp from a single 4-barrel-carb-equipped, 5.8-liter (358inch) pushrod engine.

Air does not enter the cylinder evenly all around the intake valve. Here better than 60 percent enters via the valve’s A half. This means the B half

is in less need of de-shrouding. Note that the flow exiting the valve is turning, thus generating swirl. Many World War II aero engines had a hemi-style combustion chamber. These typically employed valve angles of as much as 90 degrees inclusive. This accommodated the biggest valves, but it also produced a very deep chamber (half a hemisphere). Deep chambers were okay for the lowcompression ratios used for heavily supercharged engines. But they were bad for high-compression use because of the high piston dome needed. In practice, it turns out that the optimum angle for the intake valve from the bore centerline is about 18 degrees. For the exhaust, where shrouding is less important, the optimum angle is about 10 degrees. Chrysler has been synonymous with hemi engines from the 1950s and with good effect. Their latest iteration, introduced in 2003, is the 5.7 (and a 6.1 version that came later) engine. The engine is a very-well-conceived design with heads that flow every bit as well as you would expect a good design to do. So far, we have talked about geometric shrouding. It’s a good start to understanding what shrouding is. But simply applying it, without further thought as to other factors in the combustion chamber, is not the way to go. If air entered all around the valve in a uniform fashion, then geometrically un-shrouding the valve would work every time. However, air is heavy stuff and tends to want to flow in a straight line. There is no point in un-shrouding a portion of the valve’s circumference if there is minimal flow there, so we must first understand where the un-shrouding needs to be done to make the most use of material removed from the combustion chamber. This can be important; every CC carved out means less compression ratio potential is available.

The intake (top) and exhaust (lower) ports of the cylinder head shown here were the subject of an intense flow-bench development program. The highefficiency figures produced made it possible to develop more than 100 hp per liter from a 10.5:1 CR engine, running service-station 92-octane premium fuel. The ground rules that got this head design to the level of producing such good performance are well within the grasp of every performance enthusiast.

Ports Now that we are clear that everything in terms of flow starts at the valve seat, let’s move on to the subject of the ports themselves. Although a flow bench is a prime requirement for optimal results, there are many ground rules that can be applied to improve a typical stock cylinder head, be it a two- or four-valve design. Let’s start with the basic requirements of a port. Assuming the port is supplying a good valve-seat design, its first priority is to flow as much air as possible. The second general rule here is that whatever airflow is achieved must be done so by high-flow efficiencies, as the port must not have an overly large cross section. If the cross section is too large, the port velocity is low, cylinder ramming decreases, and flow reversion increases. The result is poor low-speed output with possibly no benefits at the top end either.

Apart from total airflow, a successful port design must also address issues relating to cross-sectional area, total length, and wet flow. Another factor that can greatly affect low- and, to a lesser extent, highspeed output is swirl- or otherwise-induced mixture motion within the closed cylinder. In addition, it helps to make sure no big problems exist with fuel management within the port. This comes under the heading of “wet flow.”

6-10. Assuming this as our starting point, we look at progressively changing the form into a much more efficient shape.

Finally, we have to look at the combustion chamber form as defined by both the combustion chamber and the piston crown. That, in simple terms, means getting the required CR without producing a poor-burning chamber in the process.

Port Evolution The most basic port we could have is a round one that has a bend in it to accommodate the valve. Using this as a starting point, we can develop a port using some simple logic. Illustration 6-10 shows our starting point. With any port, the radius of the short side turn is usually the number-one obstacle to achieving good mid-to high-lift flow figures. Formula 1 engines have a very large short-side turn radius and the port’s downdraft angle is less than 30 degrees off vertical. This makes for a very simple port that requires very little in the way of Band-Aid fixes to make it work extremely well. Unfortunately, the dictates of less-than-ultimate power on a less-thanultimate budget and the low hood lines of street vehicles mean ports that are, without doubt, severely compromised. Illustration 6-11 shows the basic steps that were taken to make the port more efficient.

6-11. The first move to make our basic port more effective is to raise the floor, starting just before the short-side turn. This allows the radius of the short-side turn to be increased. Doing this, though, leaves the port with a small cross-sectional area around the turn. To compensate, the port needs to be made wider (as viewed from above). This works fine for a Hemi or with four inclined valves as per most four-valve designs. For a parallelvalve (or nearly so) 2-valve head, the expansion of the port almost always needs to be mostly on one side—the cylinder-wall side. By biasing the port we allow, at high lift, the port to follow a form that more closely represents the direction the air wants to travel in. If you are modifying a typical parallel-valve two-valve head, then the last step is an important one to keep in mind. Understanding that the port more than likely needs a bias is the key to getting those big high-lift flow numbers. This leads ultimately to high-lift flow efficiency figures that exceed those delivered by a four-valve head. The recovering flow efficiency of a parallel two-valve head is one of the reasons why this type of head responds to high valve lift, so bear that in mind when it comes time to spec-out the cam and valvetrain. The lift to

shoot for is 0.30D for a hot street machine and as much as 0.35D for an allout racer.

Above about 0.27D, the flow of a typical two-valve head starts to move predominantly across the back of the valve.

Past 0.10D, two valves show better efficiency. But at about 0.27D, the single valve recovers and ultimately wins out.

Cross-Sectional Area

The optimal cross-sectional area for a given size of cylinder can float around somewhat, depending on the size of the intake valve being used, how tortuous the basic port is, and how big the cylinder may be. A good starting point for the intake is to have a nearly parallel section of the port, about 1.5 to 2 intake valve diameters up from the intake valve, sized to a cross section equal to 77 to 80 percent of the area of the intake valve itself. This area needs to extend, if necessary, into the intake manifold for about 2 to 3 times the diameter of the intake valve for a two-valve head and about 4 to 5 if we are dealing with a four-valve head. For the record, it is better for the widest powerband and best torque if one errs on the smaller side; this produces a punchier driving experience. Making the port too big can hurt output everywhere. Chart 6-12 shows the test results run on a small-block Chevy.

Applied Basic Porting Now that we have plowed through the basics in theory, you might be asking, “What’s it worth in terms of power?” This obviously varies from head to head and engine to engine. The better the head is to start with, the less difference your efforts from a basic porting job will be. Let’s use a middle-of-the-road example, a 5.0 Ford Mustang V-8. The results shown in Charts 6-13 and 6-14 can be achieved if we take those heads and do nothing other than skinny-down the guide bosses, so they are only about 1/8 to 3/16 inch wider than the guide bore, and round off the short side turn. All this can be achieved in just a weekend of work. If you already have a die grinder and suitable carbides for all the other jobs that these tools come in handy for, your cost is virtually zero.

6-12. The trends with increasing port size are easy to see from these tests. The smallest port (at 180 cc) produced the best torque up to about 3,300 rpm. The 200-cc port lost a little at the low end but gained a worthwhile amount at the top end. At this point, the port size for optimal results over a wide rev range were close to being achieved. The 215-cc port delivered a slightly better top end than the 200 but lost out everywhere else. The 230cc port produced no more top end than the 215 but lost in every other area.

Here are the flow curves produced by progressively enlarging the port (as measured in CC of port volume). Note the bigger ports only showed an improvement in the higher-lift ranges, showing once again that the valve and seat control flow at lower lifts.

Here’s a typical factory head casting of the 1970 to 1990 era. An abrupt short-side turn and an intrusive guide boss kills flow.

Just how bad a typical pushrod two-valve intake port (as cast by the factory) is can be seen here. The top port mold is a factory stock 5.0 intake port. Its shortcomings can easily be seen. The center port is an as-cast aftermarket head and is much better in all respects. The bottom port mold is the center one after porting. Note the seat-to-port blend.

Wet-Flow Testing Regular or dry-flow testing is something most performance enthusiasts have at least heard of, but maybe not wet-flow testing. The purpose of wetflow testing is to establish that the fuel and air arriving at the cylinder is still

suitably well mixed. In technical terms, this is called mixture quality. If the fuel separates from the air more than a minimal amount, the power takes a dive and fuel consumption climbs.

6-13. Here is the improvement seen by applying some basic porting moves to a 5.0 Mustang head. Wet-flow testing is not so much about measuring what is happening. It’s about observing and using judgment and experience to ascertain just how good (or not) the mixture quality arriving at and entering into the cylinder is. Wet-flow testing is still a science/art practiced by few, but the few who do it often see some relatively dramatic returns on time invested. My own experience investigating mixture quality arriving at the cylinders in the 1970s proved its worth by netting a couple of major championship wins. It can be a complex subject, and worth far more than the few pages I devote to it here. The preparation for a quality mixture starts at the fuel injector or the carb booster. At part throttle, the manifold vacuum can almost completely vaporize the fuel. That’s good for combustibility and is a key ingredient to making good mileage. However, at WOT, there is not much vacuum to assist with mixture preparation. The fact is that, for the most part, the situation deteriorates a short distance downstream of the injector or carb boosters.

6-14. Power gains with the basic porting job applied to the heads on the 5.0 Mustang test engine were as shown here. Peak torque is up by 10 ft-lbs and peak horsepower by 21. Our initial move here is to make sure we are using an injector or booster that does a good job. Injectors that are working properly are usually close to as good as they can be. The same cannot be said of carb boosters; so if mixture preparation appears to be a problem, investigate higher-atomization boosters. Having to run a richer-than-expected mixture to get maximum power is an indicator here, along with excessive fuel wash, exhaust smoke, and poor WOT fuel consumption. Steps must be taken to remedy the situation. The wet-flow bench points the way, but most fixes are less than obvious. But there are some moves that can be readily appreciated. The intake port “shear dam” is a prime example of this and has been very successfully used in several highly developed heads.

The Dart big-block Chevy head produced good results before wet-flow testing. The same applies to Dart’s small-block Chevy heads, which, when tested at a Charlotte test facility, showed an increase of some 12 ft-lbs and 21 hp. A lot of what it takes to get good mixture quality at the time the plug fires is about velocity. Good velocity through the port is the place where it all starts. Having good port velocity allows us to suitably direct that velocity to generate swirl (in the case of a two-valve engine) or tumble (in the case of a four-valve). Also, mixture motion produced by the quench of squish action is key. Burning the charge faster makes for not only more effective combustion but also reduces the chance of detonation. If you need to lower the compression of an engine to have it compatible with a supercharger, do not just park the piston farther down the bore. In most instances, this serves to make the engine more prone to detonation because the charge burns slower. In so doing, it has more time to radiate heat to the as-yet-unburned part of the charge. At some point the combination of heat and pressure causes it to detonate. Having a functional quench action not only allows the engine, in most cases, to make more power, but also to do it on less octane. So far we have talked of many ailments that an intake port can have in terms of its ability to negatively impact mixture quality. Along with this, the only fixes shown have been complex and require a lot of high-dollar test equipment. This begs the question as to what the guy who has a lot less in the way of flow and dyno equipment can do to make the best of the situation. Essentially, the simplest advice here is to start with a carb or injector system

that delivers a suitably fine spray and feed this into ports that are neither too big nor too shiny. A good 60-grit emery finish in the ports and intake manifold runners can do much to stave off poor mixture quality.

Compression Ratio Compression ratio is one of those commonly known engine factors that has big power implications, but is rarely understood. It has much influence on the so-called optimal engine component combination. So that we can better understand the issues involved, the first move is to establish exactly what the compression ratio is. Illustration 6-15 does just that.

Here are the wet-flow patterns produced by a Dart big-block Chevy head at 0.100-, 0.300- and 0.500-inch intake valve lift. At each increment, the vacuum was adjusted to the value seen in a running engine. The ellipse of the intake valve can be just seen in the combustion chamber top left. The blue is streaming fuel. You can see that the charge gets wetter as the valve lift increases.

This simple example shows how it works. Let us assume the cylinder has a volume of 100 cc and the combustion chamber volume (shown in red on the right hand cylinder) has 10 cc of volume. The CR is then the volume above the piston in the left-hand cylinder divided by the volume above the piston in the right-hand cylinder. The left-hand cylinder has 100 + 10 cc, while the right hand has just 10 cc. That equals 110 divided by 10—giving an answer of 11:1.

This shot closely simulates a big-block Chevy induction cycle, without actually being one. The arrow indicates the curvature of the valve. The wide, red band next to the edge of the valve is well-atomized mixture entering and moving toward the center of the cylinder. The stream from top to bottom is wet-fuel flow. Each minute dot is a fuel droplet.

This view down the intake port shows the form of the shear dam (above and to the left of the guide boss). To arrive at a shape that does not impede airflow to any extent, and shears the fuel in the manner required, can use up a lot of R&D time. Taking the shear dam out of this Ultra Pro Machining 7-liter Chevrolet Corvette port actually reduced flow.

Here the fuel-distribution pattern can be seen in a typical small-block Chevrolet intake port. The blue indicates the presence of fuel. Note the vortex on the top of the port just before the long-side turn. We can also see that the main stream of fuel is in the high-speed area at the short-side turn. From here, the fuel stream mostly skips across the back of the valve and enters on the long-side turn of the seat. Almost no fuel is in the longside turn. After entering the cylinder, most of this fuel impinges on the

combustion chamber wall adjacent to the spark plug and is seen as fuel wash.

The shear dam to the right of the intake guide boss can be clearly seen here. A fuel rivulet arriving at the dam is sheared off and re-introduced into the air. From here, it mostly exits the valve in the six- to seven-o’clock position and, in doing so, aids swirl production. After crossing the port dam, remaining wet flow is now sheared by the ridge along the spark-plug boss.

6-15. The compression ratio is the ratio of all the space above the piston when it is at the bottom of the stroke compared to all the space above it at the top of the stroke. As the CR goes up, there is also an increase in torque output. Theoretically, this increase should occur throughout the RPM range, so power goes up correspondingly. Illustration 6-16 shows how output increases as the ratio from one point to another is increased. We glibly talk of the compression ratio and understand that it is a function of the piston traveling up the bore on the engine’s compression stroke. We need to appreciate just how significant the compression ratio is to our ability to spec-out an optimal engine parts combination. And therefore we must consider what happens as the piston comes down from top dead center (TDC), on the power stroke, in terms of the expansion ratio. High-Compression Chambers

The fact that increasing the CR is good for output does not give us a license to just dive in and crank that number up without giving due consideration to the consequences. As the CR is increased, so the heat of compression and combustion temperatures increases. This makes it harder for the air/fuel mix to resist spontaneous ignition (detonation). When detonation occurs, the resultant super-rapid pressure rises and temperatures rise to where they quickly destroy pistons. Just how much compression can be used depends on many factors, with the most commonly appreciated one being the fuel’s octane value—the higher the octane, the greater the fuel’s resistance to detonation.

An engine’s thermal efficiency, that is, the efficiency with which it can turn heat energy into mechanical energy, is directly related to the compression ratio. To better understand this, we need to look at the engine’s expansion ratio, which in effect is the other side of the coin to the compression ratio.

6-16. This chart shows how power output rises with an increase in CR. An example looks like this: Assume our engine currently has a 9:1 CR (9/1 in the column on the left) and we wish to increase it to 12:1 (12/1 in the row across the top). Where the two intersect is the percent gain in output from the increase. In this instance it’s 4.5 percent. Regardless of the fuel’s octane value, we should strive to achieve a combustion cavity that is as resistant to detonation as possible. This means producing a compact shape that has good mixture motion within and so burns rapidly. It is easy to get carried away with thoughts of a big compression number and overlook the chamber that may result. Keep in mind that a charge trapped in the quench area does not burn well, if at all, until the gap between piston and head exceeds about 0.100 inch. Also, big piston domes do not help achieve an effective burn. From personal experience, I can say that a piston crown shape that inhibits rapid and complete combustion can wipe out 10 percent of the engine’s potential output. When building an engine, it is best to machine the head to minimize the combustion-chamber volume first, and then use a piston with the appropriate crown shape to get the desired CR. In almost all cases, shallow-dish or flattop pistons are best. However, a well-thought-out raised crown for those higher ratios (12:1 on up) can pay off big time, but it may take dyno testing to come up with what is best for the job.

A piston crown with an edge shape as indicated by the arrow can severely inhibit the passage of the flame through the mixture. To make the most of the CR being used, the flame front must progress rapidly and in an unobstructed manner through the charge.

CR—How High Can We Go? If we put an engine on the dyno and keep cranking up the CR, the gains produced taper off until, at about 17:1, no further gains are seen. The increase from about 14:1 to 16:1 is very minimal, so why would we want to go up to these super-high ratios in the first place? This is a good question, and it is a classic case of getting the right spec/parts combination that ultimately produces the winning results. The CR affects several important and power-influencing factors. Understanding these will put you more than one step ahead of the competition, so let’s investigate. The two most influential factors the CR affects are: the balance of the intake-to-exhaust-valve size, and just how much cam the engine can successfully use. It is important to understand that the valves for a high-performance engine need to be as big as possible, bearing in mind any mechanical constraints. For a two-valve engine, this results in valve proportions in relation to the bore as per Illustration 6-16.5. Having established that cramming in the most valve area possible is a requirement, we have to ask ourselves how much we should apportion to each valve. The usually touted figure here is that the exhaust valve (assuming the same intake-to-exhaust flow efficiency) should be 75 percent of the intake valve. Although some engine builders stick to this number as if it were immutable, the fact is it floats in relation to the CR used. To see how this works, we need to move from considering the compression ratio to considering the expansion ratio.

In-cylinder pressure measurement can tell us a lot about an engine that is useful for finding yet more power. Here the uppermost curve is the combustion temperature in degrees C. The blue curve is the cylinder pressure. In this instance, it peaks at 1,536 psi. Note the spikes in the curve near the peak. This is the result of an inconsistent burn and/or a chemical reaction, such as the generation of oxides of nitrogen (NOX), which absorb a portion of the combustion heat. The lower curve is the cylinder pressure that would have occurred if the plug had not fired the charge.

6-16.5. Here are the typical valve proportions for a high-performance twovalve engine in terms of bore size. For engines with the CR in the 10 to 12:1 range, the diameter of the exhaust valve is typically 75 to 80 percent of the intake, but this changes with CR. Take a good look at Illustration 6-17. First, let us deal with how the CR (and by inference, the expansion ratio) affects the valve-size proportions. Let us assume two cylinders—one with a 15:1 CR and the other with a 2:1 CR. While they are at TDC, let us fill both cylinders with compressed air to 1,000 psi. If we now let the pistons go down the bore, the pressure for the 2:1 cylinder decays as per the darker of the blue lines, and the pressure for the 15:1 decays as per the red line. Note how the pressure drops far faster for the higher compression. Next, let us say that we are going to open the exhaust valve 90 degrees before bottom dead center (BBDC). Doing so with the 15:1 CR cylinder means that the pressure has dropped to about 18 percent of the original; so 92 percent of the pressure energy has been used to turn the crank. If the exhaust valve is opened at the same point of the 2:1 cylinder, the amount of pressure energy remaining and subsequently wasted is amost 50 percent.

6-17. Although there are only three curves on this graph, the relationship between CR and cylinder-pressure decay is more influential than it may first appear. The inference here is that a high-compression cylinder does most of its useful work earlier in the expansion cycle. Given a high enough CR, opening the exhaust valve earlier than normal has little affect on the power produced by the cylinder. On the other hand, opening the valve early on a lowcompression cylinder throws far more pressure energy away. For a highcompression cylinder, the consequences are that the exhaust valve can be opened sooner and therefore does not need to be as big. Making it smaller means more room for the intake valve; so the ratio of exhaust to intake for, say, a 17:1 Pro Stocker, is going to be about 65 percent. Conversely, a low-compression engine needs to hold the exhaust valve on the seat longer to make the most of the cylinder pressure. This means it has to be bigger so that, when opened, it has blown down sufficiently to minimize exhaust stroke pumping losses. Here, we can be talking of an exhaust-to-intake ratio of 80 to as much as 85 percent. Because of its inherently lower CR, a supercharged engine needs to have a bigger exhaust in relation to the intake than does a non-supercharged engine. The fact that a high-compression cylinder does its work earlier in the cycle also means that a shorter rod can be used without such a big sideloading friction penalty. You can see how the CR affects the optimal rod length and the ratio of intake to exhaust. There are few one-size-fits-all parts

or specifications within an optimal high-performance engine. This explanation of how the CR affects valve sizes and rod lengths for best output also demonstrates why changing just one thing on the dyno to see if it works is not necessarily the thing to do. It also demonstrates why assuming one aspect needs to be a certain value, come what may, can be a mistake. In practice, a 2:1 cylinder does not achieve peak pressures as high as those seen by a 15:1 cylinder. Peak pressures increase as the CR goes up and a good general rule here is that, for a street engine, the peak is 100 times that of the ratio involved (i.e., 10:1 produces 1,000 psi). For a race engine, where volumetric efficiencies can exceed 100 percent, the peak pressures can be as much as 120 times that of the CR. Again, take a look at Illustration 6-17 and compare the red line with the light-blue line. These, on the same scale, represent the starting and finishing pressures of a 2:1 cylinder compared to a 15:1 cylinder. The shaded green area between the two shows the difference in the available pressure to turn the crank. Cam Specs and Dynamic CR One way of boosting power is to throw in a long-duration cam, but the down side is that such a cam favors top-end output at the expense of low-end output. Worse yet, it affects the CR while the engine is running. This is known as the dynamic CR, as opposed to the static CR discussed above. If the cam opened and closed the valves at TDC and BDC, then both these CRs would be the same. However, valve events are spread such that the valves open sooner and close later than the TDC and BDC positions.

Drawn here to scale are the relative positions of the piston in relation to the intake duration of various cams. As you can see, the installation of a 300-degree race cam seriously degrades the dynamic compression ratio, and consequently the cranking pressure seen, unless the static CR is increased accordingly.

6-18. Increased compression can be a vital ingredient for making a longerduration cam deliver its best. An appropriate CR increase with a longer cam works wonders for the entire curve. For this test a 265-degree cam (gray curve) was substituted for a 285-degree cam (blue curve) and, without a change in CR, resulted in a substantial drop in low-speed output. Raising the CR from 9:1 to 12:1 (green curves) recovered almost all the lost low-end output and gave a further big increase in top-end output. Since intake valve closure is delayed, the piston has moved up the bore somewhat before the valve closes and traps the charge. This means that, at low speed, the CR that is actually seen by the cylinder is somewhat lower than the calculated static CR. If this results in too low of a CR, low-speed output suffers accordingly. Initially, a certain amount of intake-valve-closure

delay (about 30 degrees after BDC) has little effect on the dynamic compression because the crank/rod geometry is such that initially the piston moves very little up the bore, but that state of affairs does not continue for long. A good working rule here is that (assuming the cylinders seal up as they should) the cranking pressure for a street engine running premium fuel is about 190 to 200 psi. If the engine is below that, then there are ft-lbs still to be had, and above—you might want to use an octane booster of some sort. Big cams need higher compression to not only allow the production of the best torque and top-end output but also low-speed output. Choosing a cam that is too long for the CR is usually a good way to lose low speed and gain little or nothing at the top end. Being a little more conservative on a cam selection can actually result in more power everywhere. There are other factors involved with the compression ratio used and the cam spec for optimal output, and these are discussed in Chapter 10.

CHAPTER 7

PORTING AND FLOW TESTING As I have established, high-airflow potential, especially through the cylinder head, is a key element in producing power. If you are starting with as-cast parts, a means of reshaping them and testing the subsequent flow results are very useful aids to maximizing output. In short, the subjects in this chapter are physically porting heads, intakes, etc.; what equipment you might need; and the results you may reasonably expect. So far we have talked about flow and flow bench results. Let’s be clear about one thing: It is entirely possible to exist without a flow bench to verify what you have done. But doing really good porting work is in the details, not the overall. Sometimes what seems like an obvious move to increase flow does the opposite, and what could possibly have been a big overall improvement is reduced to something a little less. Two or three mistakes like this may mean the difference between winning and being a runner-up. A flow bench is not only a real aid to finding power; it’s also fun to use. At this point, you may be convinced you need a bench but the possible cost and learning how to use it may seem too daunting. Well, the good news is you do not have to spend much money here. Instead of buying a $10,000 pro-built bench, I am going to show you how to do the job for about 2 to 3 cents on the dollar.

From the upper view of this Ford 5.0 Mustang head, it is easy to conclude that the guide bosses are hardly conducive to good flow. Doing nothing more than the seemingly obvious of streamlining the guide bosses (as in the lower photo) increased rear-wheel output of a stock machine from 198 to 214 hp. The time involved was less than three hours.

Building a Really Trick Low-Cost Bench May 2008 was a Vizard milestone as far as modifying engines goes. It represented, for me, 50 years of cylinder head porting and flow testing. Initially, the methods I used were crude and simple to the extreme, but what I did worked very well and produced top-notch results. Looking back on these early efforts, I realized many years later I should have continued flowing heads using my original method, instead of being swayed by convention. Maybe you’re already familiar with the use of a flow bench. If so, you see that, once again, I do fly into the face of convention—but not without good reason. Here, starting back in 1958, is how events unfolded. My first flow bench was in fact my mother’s vacuum cleaner. I mounted the cylinder head to be

tested (an A-series head as per Austin A35 and later the Mini Cooper engine) on the block with a means of opening the valves. I then attached the suction side of the vacuum cleaner to the bore and sealed the hose so there were no leakages at this point. Next a spark plug with the middle removed and a piece of 1/4-inch-diameter copper tube glued in was installed into the spark plug hole and connected up to a manometer made from clear plastic tubing stapled onto an 8-foot-long piece of 2 × 4 wood. This was marked out in inches from 48 inches at the bottom all the way through zero to plus 48 inches at the top (note: a more powerful vacuum cleaner needs a taller manometer). The plastic tubing was looped so that the bottom of the “U” was formed about a couple of feet below the bottom of the 2 × 4. Water with food dye was used to fill the “U” section of the U-tube until it reached the zero mark (see Illustration 7-1). As you can see, there is not much to this bench. It can be constructed in a few hours for a minimal expenditure. If you already have a shop vac, then the rest of it probably costs less than thirty bucks. How It Works Here is how this simple bench works: the more open the valve is, the lower the pressure drop as measured by the manometer. At low valve lift, let’s say 0.050 inch, the manometer can read (depending on the vacuum cleaner’s capability) anywhere between 60 and about 100 inches of water. If the flow at a given lift is improved, the manometer reading at that lift drops. Let’s say that the stock head at 0.050 valve lift produced a reading of 80 inches. We then do some seat blending on the valve and head casting and on the next test the manometer reading is 70 inches. This means the flow has gone up, so the pressure drop pulled by the vacuum cleaner for a first approximation, the flow has increased by the square root of 80 divided by 70. That works out to 1.069, or 6.9-percent improvement.

7-1. The type of flow bench shown here is the floating-pressure type. Round up the minimal parts required, and you can build one like this on a Saturday morning. If you already have a block and a vacuum cleaner, the rest of what is needed can usually be sourced for less than $75. At this stage, we have a bench that shows only whether the port is better or worse. Now, here comes the aspect that separates the methodology of use with this bench from 99.99 percent of all others out there: By measuring the pressure drop to determine “better-or-worse” scenarios, we can see that it is not testing the head at a standard pressure drop as per almost every pro-built bench does. The standard-pressure-drop-measuring method applies a fixed amount of suction/pressure to the head as it is tested. For instance, a big Superflow bench typically uses a 28-inch pressure difference across the cylinder head.

As the valve is opened, the suction drops; so the operator opens up a control valve to bring the applied pressure back up to 28 inches. To determine whether or not flow has improved requires that the results have to be calculated and converted to cubic feet per minute (CFM). This is the complex, math filled, part of flow bench design that deters most people from building their own bench. At this point, you may be thinking that not having the ability to establish the head’s flow in CFM is a small price to be paid. The most important feature—that is, determining better or worse—has been achieved. At the end of the day, that is what really counts. After having built my “bench,” I spent a couple of years wondering how I could get real CFM numbers off the bench like the guys at Westlake (at the time the big, if not biggest, name worldwide in airflow). Then one day I was talking to an engineer who was somewhat older than I and he happened, during our conversation, to drop the key words for me. These words were “standard pressure drop.” Well there we go; I then realized that if I could adjust the test pressure/vacuum so that it was a constant, then I could, through a big bunch of complex math, calculate the CFM. So over the next few years, I built increasingly more sophisticated benches until finally, during the early 1970s, I built a bench that conformed to British Standards as far as precision gas flow measurement was concerned. It was a highly accurate 15-foot-long monster. Cylinder head guys came from 200 miles around to get their work tested. I even did some work on it for the McLaren F1 team. Some Rethinking By the mid 1980s, I began to have second thoughts on the use of a standard pressure drop to test heads. The issue here is the sort of pressure differences seen between the port and cylinder in a live-running race engine. In 1984, S-A Design published Power Secrets by the unforgettable Smokey Yunick. The author goes to some lengths to show that the pressure differential needs to be above a certain value to get dyno results that sort of coincide with what you might expect from any flow increase. The number that Smokey came up with was 28 inches of water. That is about an industry standard as of 2010. But here is the problem: A race engine does not see a fixed standard pressure drop. Here, we need to take a look at what actually happens.

Although referred to as a four-cycle engine, a high-performance/race engine is actually a five-cycle engine. The fifth cycle (No. 5) is the exhaust-driven start to what subsequently becomes the regular pistondriven induction stroke (No. 1). The induction system on a true race engine is, for the most part, exhaust driven. That is, the scavenging pulse from the tuned exhaust pulls a far bigger depression in the cylinder than the piston going down the bore. On something like a NASCAR Cup Car engine, this can amount to as much as 120 or more inches at TDC in the overlap period. The draw on the intake port at TDC can be so strong that, even though the piston is virtually parked, the intake charge in the port can be traveling into the cylinder at as much as 90 mph! When the valve is near wide open and the piston is moving at peak piston speed (this is between 72 and 74 degrees after TDC), the draw on the valve is between 15 and 20 inches of water. All the forgoing leads us to one conclusion: If we want to more nearly simulate what happens in a running engine, our intake flow testing needs to be done at a high-pressure drop at low valve lift, and a lower one at high lift. This is exactly the way the flow test rig illustrated here functions. An

uncontrolled vacuum source (such as a shop vac) pulls a large vacuum when the intake valve is closed and a progressively lesser vacuum as the intake is opened. So running a “floating” pressure drop, as we are doing here, is a more realistic simulation of actual flow during engine operation.

Within the induction of an effectively exhaust-tuned engine, the negativeexhaust pulse pulls a far higher vacuum on the intake port than does the piston traveling down the bore. This high vacuum occurs at low lift, thus justifying the use of a higher flow-bench pressure drop at low lift and a lower one at high lift. The flow bench produces a floating-pressure drop.

Shown here is the port velocity of a two-valve race engine just prior to peak torque RPM. Note the intake velocity at TDC at the start of the intake is 85 feet per second. That is almost 60 mph! At this point, we have a flow situation that more closely mimics the pressure differentials seen in the cylinder/intake port of a running high-

performance engine. So what are the advantages of testing in this manner? Let’s analyze that now. If the pressure drop used is too low, the flow pattern that develops is not the same as it would be at a higher pressure drop. If we use a low-enough pressure drop, the flow will be virtually laminar. Fortunately for us, laminar flow becomes unstable and turns to a turbulent flow at pressure differentials well below whatever we are likely to see or use. Basically, laminar flow never occurs in a running engine; the flow is always turbulent. Returning to the effects of low-test pressure differentials, let’s say only 2 inches is used. At such a low-pressure drop, the flow in the port is slow and air stays attached even around the worst short-side turn. When a result from a test like this is corrected to, say, the commonly used 28 inches, it produces numbers that are much higher than if the head were really tested at 28 inches. By running tests at real-world test pressure drops, we create the same pattern of flow-reducing separations as occur in actual use. Here’s an example to illustrate the situation: Take a typical aftermarket highperformance 23-degree small-block Chevy head and flow test the exhaust port at 28 inches. At this test pressure, it does not lag that far behind a megabuck NACAR Cup Car head having the same size of valve. Up to about 0.500-inch lift, the 23-degree head nearly mirrors the Cup Car head. At a little over 0.500, the Cup Car head slowly shows its superior capability. At 0.700 lift, the Cup Car head has about a 15-percent advantage over the best 23-degree heads. Step up the test pressure to 120 inches, and the picture changes dramatically. At about 0.200 inch, the flow on the 23-degree head separates on the short side. This partially shuts off part of the port area that is normally used for the high-speed flow on the long side. With the Cup Car head, there is no real flow separation until about 0.500 lift, and what does occur is relatively minor in nature. The consequences here are that, from 0.200 on up, the Cup Cars port pulls away from the 23-degree port. And by the time 0.700 lift is reached, it is close to a 20- to 25-percent improvement. From this, we conclude that port modifications that produce the most positive results are doing so by addressing those flow patterns and improving the port shape to best deal with them. The bottom line is this: A cheapo flow-bench test setup is actually a better tool for developing an intake port than a $10,000 commercial flow bench. The only down side at

this point is knowing just how many CFM the head is flowing if each reading could be corrected to the common 28-inch pressure drop. Without this number, you won’t be able to make a comparison with other flow test results to be able to gauge your work compared to others. If you really need results in CFM, this can be fixed relatively easily as we shall see; but for now let’s consider the exhaust.

Flowing the Exhaust Without making some rather fancy test equipment, we are not going to be able to flow the exhaust at real, live test pressures. Normally, when an exhaust valve opens, the cylinder pressure is somewhere between 90 and 120 psi. If you are intent on having a pump that develops this kind of test pressure, even for just the low-lift tests, then be aware that you need about a 200-hp motor to drive the pump. Very few flow bench setups are capable of this (though I understand from some of Ford’s engineers that their bench can approach real-world exhaust pressure drops), and that it costs a mere sevenfigure number to build. For the most part, we flow the exhaust at 28 inches, and live with the fact that it is not the best way to do things. However, we find once again that our low-buck bench, with its uncontrolled floating pressure drop, actually does a better job than a commercial bench at a fixed pressure drop. When I first wrote about this bench in my online magazine, flow bench guys bombarded me with emails, saying it would not work, could not produce accurate results, etc. It seemed so ridiculous for someone to make a comment that something can’t be done after it has been proven. And as for accurate results, given the application of a few precautions, this bench design can produce extremely accurate results and repeatability. Establishing Repeatability and Accuracy The most critical factor in getting usable data from a floating-testpressure flow bench is making sure that the reading you take today will, if nothing changes on the test item, be the same tomorrow, next week, and next year. The simple way to do this is to have two or three test orifices and check that each delivers the same pressure drop it originally did when tested. If it does not, then the voltage to the motor has changed, and we know that happens just from demand from the rest of the circuit community around

you. The way I stabilized voltage, back in the day, was to use a transformer to step up mains voltage by about 15 percent and then drop it to a standard voltage with a rheostat. That way, the motor voltage could be finely adjusted so that the test orifices showed the pressure drop originally assigned to them. Quantifying Results in CFM My friend Roger “Dr. Air” Helgesen built a bench that worked along the same lines as this in the late 1970s and still uses it to this day to flow heads and intake manifolds. As usual, Roger adopted a singularly simple way to convert the pressure drop seen on the manometer to CFM at 28 inches, using nothing more than a sheet of graph paper and a few calibration orifices. Accuracy on the order of 3 percent (that’s about the same as a typical, commercially available bench) can be had here if suitable care is taken to eliminate unwanted leakages, stabilize supply voltage, and make frequent reference checks against the calibration plates.

7-2. These holes need to be machined to a very smooth finish. As for tolerances, make them as accurate as possible (+/- 0.001 is acceptable). The hole sizes and X–Y coordinates are: Hole 5: 0.210 Dia. on 4.30/3.40 inches. Hole 10: 0.296 Dia. on 3.95/4.10 inches. Hole 20: 0.419 Dia. on 3.05/4.30 inches. Hole 40: 0.594 Dia. on 2.10/3.80. Hole 80: 0.840 Dia. on 2.05/2.5 inches. Hole 160: 1.185 Dia. on 3.70/2.40 inches. Radius entry for all holes is 0.25 inch. This is machined such that the edge of the radius goes out to 80 degrees, not the full 90. The back of the 5- and 10-size holes must be chamfered with a 90-degree cutter. For the 5 hole, the chamfer should be 0.460 diameter, and, for the 10 hole, 0.546 diameter.

7-3. By plotting the individual orifices against the depression seen, a graph can be developed. The subsequent pressure drop with the head being tested can then be converted to CFM.

It is entirely practical to flow test a head using no more of a vacuum/pressure source than is delivered by a typical shop vac. Illustration 7-2 shows dimensions for these calibration orifices. These flow a certain amount of air when tested at 28 inches. By drawing a graph as per Chart 7-3, a curve can be developed of depression versus flow at 28 inches for your particular setup. Build a bench like this and you are in business.

Porting Follow what I have to say here and your home porting costs will be a lot less than just going down to your local industrial tool-and-cutter supply outlet. Up to now the subjects covered have looked at what is needed in the way of port and combustion chamber design to make the most of any particular head casting. What we now look at is the cutting, grinding, and finishing processes that are needed to produce whatever port and chambers are being reworked.

If you want to upgrade a bench, it can be done for a lot less than you may think. Here the Audie Technology Flow Quick setup is being tested. Essentially, this is a computerized, instrumented version of the floatingpressure bench detailed earlier. For full details on this bench go to www.motortecmagazine.com Die Grinders If you have no porting equipment already at hand, then a decision is needed on two points: cheap or expensive, and electric or air. My advice here is start cheap. As to air or electric, you need to make that decision on your own. So we need to port like pros, yet do it on a typical hot rodder’s budget and that leads us to the next question: Do you already have air in your shop? If not, do you plan on getting an air compressor for other needs such as blowing off parts, running air tools, using a spray gun, etc.? If your shop already has air and if you have to buy a die grinder (bearing in mind that 18,000 to 20,000 rpm are needed), the best route to go is with an air-powered die grinder. The principle reason for this is that there are some really cheap die grinders that can be had for as little as $20. A decent electric die grinder that lasts any reasonable time costs five to seven times that amount. The

cheaper air grinders aren’t as well made, but I have a fix that extends their life from 3 months to 3-plus years.

These days, you do not need to spend a fortune on compressors to drive an air die grinder. I bought this one in 2005 for $285, and it meets all my grinding/porting needs. Even if your shop does not currently have compressed air, I recommend you seriously consider the air grinder route here. If you are building engines, you need shop air for many reasons. There is an old saying, “You only get what you pay for.” That holds good for cheap die grinders. Cheap grinders use even cheaper bearings, seals, and the like, and as a result wear out a whole lot faster. My experience with a substantial number of cheap die grinders is that most only do about four to six pairs of heads before becoming nearly unusable. However, I have found that, of all the different air tool lubes I have used, the one from BND can

extend the life of a cheap grinder by a factor of ten—good for about a hundred individual heads.

Here is what I use to lube all my air tools. As of 2010, it is the best I have found. And for the record, Boeing seems to agree with this finding. With cheap die grinders, you don’t have too many choices as to the body style of the die grinder. More expensive die grinders are offered in several styles that can be categorized basically as short- or long-body. In the cheaper range, it is almost inevitably just the short style that is offered. This is okay. The long-body grinders do offer an advantage when it comes to porting intake manifolds but, for 95 percent of the uses you are likely to encounter, you can get by with a short-body grinder. Having acquired a die grinder that goes fast enough, then you have to think about how to slow it down; not all jobs call for a speed upward of 18,000 rpm. Like die grinders, a speed control need not cost that much. A simple air-pressure regulator gets the job done.

After three years of extensive use, my previous cheap die grinder was replaced with this cheap die grinder. BND lube is the key to long life from cheap grinders.

If your air compressor does not have an easy-to-use pressure regulator, then get yourself a budget one (as seen here) to control speed. Removing Metal—Fast

Porting is all about reshaping ports and combustion chambers to more effective forms. And 95 percent of the time this involves removing metal. If you are short on knowledge as to how best to do this, it can be tedious beyond belief!

For most porting jobs, a small variety of carbide cutters suffices. For aluminum, a coarse-pitch tooth form is needed, and for iron, a fine one. The two materials most likely to be reworked are aluminum and cast iron. Each, to an extent, calls for a particular specification of carbide cutter. Though initially not so cheap, a $25 to $30 carbide can easily port 20 aluminum heads and be re-sharpened for about half its initial cost. Monetarily, it looks like a straightforward deal here but the price of carbides is highly variable. The cost of a quality carbide can vary from as little as $15 a pop to as much as $35 for what is essentially the same tool.

Where you purchase the item has a lot to do with the cost, and there seems no real set rule here. Generally, I avoid professional-porting supply and tool distributors. The customers they deal with are usually machine shops employing 5 to 105 machinists who have to be kept busy or money goes out the door. But there are discount shops that, though a lot less common, can be found in bigger cities. And there are auto swap meets. Surprisingly enough, you can often get porting supplies at such events at bargain rates, but you do need to know what you are buying and what you should be buying. Let me explain. You can purchase carbides that cut a smooth finish or you can buy carbides that remove metal at a rapid rate, and sometimes you can buy a carbide with a tooth form that is pretty good in both arenas. But would you recognize one from the other? You probably would not, because it takes an experienced eye. You can see a little problem cropping up here with cost and performance. Fortunately, I have the solution: Purchase Dr. Air porting supplies from Dr. J’s Porting Supplies. The bottom line here is that, worldwide, there may be thousands of places you can get carbides, but there are probably only a handful in the entire world that can sell you, at significantly below-regular prices, carbides explicitly designed and produced for the porting community. Aluminum and cast iron have radically differing properties, so it’s hardly surprising that, though there is a little overlap, each requires its own style of tooth form for optimal results. In essence, we can say that cast iron normally requires a finer tooth pitch than aluminum. Too fine a tooth pitch and aluminum clogs up the cutter in nearly no time flat. Too coarse a pitch on cast iron and the cutter dances all over the place and barely gets to grips with the job of removing metal. As may be expected, there is a fair amount of technique and technology involved in porting. We look at a little more of what is required here but, for the most part, this subject is dealt with in detail in my porting school in online magazine at www.motortecmagazine.com.

A small, 50-cc bottle of the Dr. Air cutting fluid goes a long way. This is how much was left after porting four sets of heads. Final Shaping and Finishing Given the right die-grinder speed, cutters, and cutting fluid, ports and chambers can be reshaped to the point they need little work with abrasives to finish them off. However, for the “pro look,” a finishing process is needed. But let’s be clear here: A highly polished finish is, in most cases, not only unnecessary, but also detrimental to power output. A shine in the chambers is fine because it cuts heat transfer. In the ports, though, and especially in intake ports, a decent 60- to 80-grit finish gets the job done. Always keep in mind that it is shape that rules, and not shine!

Most cartridge rolls are supplied in boxes of 100, which is okay for a pro shop but is a little pricey for the home porter. These Dr. Air 50-count boxes are much more wallet friendly. Finishing ports off requires such things as abrasive rolls, disks, and flap wheels. These all come under the heading of consumables, and they can get expensive. It may sound like I am replaying an old record here but, once again, I suggest using Dr. Air products. Depending where you might otherwise source them, Dr. Air consumable abrasives are about 40 percent cheaper. Equally important for many end users is that they are available in Bonus Boxes of 50-plus at a time. Also from the same source, you can get a starter kit, which ports more than a set of V-8 heads.

Results If you are going to build engines, you are going to need the equipment I have described in this chapter, even if you elect not to port your own heads but buy them. Such jobs as port matching and generally cleaning up intakes, oil passages, oil return routes, and the like, call for the same tools as porting heads. If you do port your own heads, what kind of return might you expect? The answer here depends on just how good or bad the initial head casting may be. The following examples should give you at least an inkling of what might be expected.

If you are starting out on your porting ventures, here is what I recommend as a starter kit. It finishes a set of V-8 heads. Starting with a typical V-8, we can expect that pocket porting nets a useful gain. A 360-hp 350 with stock 186 factory head casting went up to 388 after a morning’s work on the heads. A mildly modified 2-liter nonVTEC Honda climbed from 155 hp at the wheels to 167 with pocket porting, and 170 with a basic full-porting job. A basic full-porting job on an MGB head, along with the intake cleaned up and some reworking of the twin 11/2-inch carbs, pushed rear-wheel horsepower from 64 to 80. Fully porting heads for two-valve applications can really pay off. A small-block Chevy that made 404 hp on stock factory 492 heads delivered 482 hp on a set of fully ported Dart 200-cc-port aluminum heads. This is what happens when intake flow goes up from 184 cfm at full valve lift to 303 cfm, and exhaust from 124 to 210 cfm. A 2.0-liter Pinto with a stock head made 130 hp and, with an all-out head, 152 hp. Heads are not the only component we can port. A clean-up and a port matching job on a small-block Chevy Edelbrock Super Victor only took about 3 hours of work, and boosted power from 556 to 571 hp on an engine of 383 cubes. Can’t afford an intake? A 10-hour porting job on a stock Q-Jet small-block Chevy intake bumped output from a near-stock 350 from 282 to 304 hp.

These flow curves are a result of the pocket porting shown earlier in the chapter. A 16-hp increase was realized for a minimal cost; plus, three hour’s works represents a good return on investment of time and money.

These days, small-block V-8 head castings from EQ, RHS, and Dart are very “porter” friendly. Even a beginner of reasonable dexterity can produce a professional-looking and -performing end product.

Thirty hours of basic porting on this small-block Chevy EQ head produced very positive results. The flow produced was sufficient to top 600 hp on a 383-ci engine with a 12:1 CR.

This Ultra Pro Machining CNC-ported LS6 Chevy head not only had more flow but the intake swirl was also increased over stock. With a minimal cut out of the intake to keep velocity up, this head improved output at low RPM while delivering more than 40 hp over stock.

Edelbrock offer this CNC-ported Victor head for the small-block Ford. My dyno tests with this head have shown stellar results, especially with stroker motors.

Here is the chamber of a race-winning four-valve head I did in about 2005. The thermal-barrier coating is plain to see. What is not so obvious is that this is a polyquad design using four different valve sizes to promote swirl and cut cross flow from intake to exhaust during the overlap.

Do you feel inclined to port your own heads? If not, CNC-ported heads from a serious, professional porting shop are an option, but at a cost. Shown here are the high-flowing 227-cc port Chevy heads from Dart.

On most four-valve heads, the port dividers are rounded at the leading edge because a cast sharp edge is not very practical to do. By knife-edging the divider as seen here, measurable extra flow can be found for minimal effort.

Knife-edging the trailing edge of the divider on the exhaust is not a good move. Target a rounded form (as seen here) for best results.

CHAPTER 8

IGNITION SYSTEMS The greater our success in terms of porting and increasing compression ratio and engine RPM, the harder it becomes to effectivly light-off the compressed charge. Fortunately, improving the ignition is not a great challenge these days and, for the most part, not that costly either. But the situation has not always been so. Until the 1970s, the contact breaker (or, as they were often called, points) system dominated. In such a system, the spark was triggered when the contacts were opened by the cam in the distributor. This system was prone to many mechanical and electrical problems. From the mechanical aspect, the points were prone to bouncing near the top end of even a stock engine’s RPM range. They also needed continual adjusting to maintain the correct gap. This was important; it controlled not only the coil charging time but also just how cleanly the points seperated on opening. These factors affected spark intensity and RPM capability. Also the contact breaker faces would electrically erode, which in turn prevented a sharp break at the moment of opening. This cut spark capability. In all, the contact breaker system was a small performance nightmare.

There are many myths surrounding ignition systems and what is needed for best results. In this chapter we cut to the simple truths of what is needed for a functional system. Back then (1967 or so), when I was racing a 9,000-rpm 1-liter Mini Cooper, a change of points and a very careful setup was done prior to every race. Not only that; we used selected points with the strongest spring we could find. That often meant going through as many as 10 sets to find a set we would actually use. With the right cam profile and a set of strong-spring contacts, we could get as many as 10,000 rpm from our 1-liter engine and still have it fire just fine. If you are restoring an early performance car, you may want to stick with the contact-style distributor. And the good news is that current aftermarket contact-style distributors incorporate just about all the go-fast moves that were ever done back when. You still need to adjust or replace the contacts regularly, but other than that, they do perform.

For our race program in the 1960s, all the fuss and bother with contact breaker ignitions was more than we needed. At that time (about 1963), I was an apprentice in the aircraft industry and studying for my engineering degree. I had access to many cutting-edge electronic components because the laboratory I worked in dealt mostly with advanced electronics for the aviation industry. One of my bosses (an incredibly smart mathematician listed as one of the top 10 in the world), suggested I make a transistor circuit to handle the switching on and off of the coil with a big-power transistor and leave the points to handle the very low current and voltage necessary to trigger the transistor in the first place. I designed a simple circuit to do just that. It worked because reliance on the contact breaker to produce a strong spark was virtually eliminated. The system could run with a very small contact breaker gap setting, and this increased the RPM capability substantially. I ran a spin test in a lathe to about 5,500 rpm (equivalent to 11,000 engine rpm) and still no sign of contact breaker bounce.

Finding a points-style (contact breaker) distributor in my arsenal of speed parts proved fruitless but MSD’s Todd Ryden saved the day. The arrow indicates the points. This early kind of transistorized ignition was known as a TAC (transistor assisted contacts) system. Soon after building this system, I found I was far from the first to do so. Indeed, some electronic whiz kids were already building systems that relied on something other than a set of mechanical springs and contacts to trigger the collapse of the coil’s magnetic field that subsequently generates the high voltage for spark production.

The common use of the contact breaker system was destined to come to an end. The only surprise here was the time it took the auto industry at large to do so, and for the electronic system to become universal (early to mid 1980s). But all that was back when. These days, ignition can plague us, but the fixes are often as easy as replacing a cheap part.

Number-One Ignition Goal Generating a big fat high-temperature spark at the spark plug, at the appropriate moment, is the number-one objective here. As hot rodders, we may have been less than enthusiastic about cars having to meet powerrobbing emission standards, but the reality here is that it forced manufacturers to drastically improve the ignition system’s performance, and that was good for us all. Modern systems are generally good to excellent, but that does not mean they cannot be improved upon. Starting at the spark plug, let’s see what we need to do to maximize performance.

Spark Plugs The first step toward making sure your ignition system is up to the job for your engine is to establish that the plug is in the right heat range. To do this, run the engine hard under full load for about 30 seconds, then shut the throttle down, and cut the ignition. Pull the plugs and primarily check the center electrode and the insulator core around it. If it shows any sign of blistering or overheating, a cooler-running plug is needed. Remember, if the plug gets too hot, it brings the onset of detonation and the likely destruction of pistons. Illustration 8-1 shows how a cooler-running plug differs from a hotter one. It is better to run with a plug that is a little too cool rather than one a little too hot. If the plug is way too cool, if there is a possibility of it fouling up; it is not reaching a high-enough temperature to burn off deposits. Plug Gap Assuming the plug temperature rating is on the mark, the next issue to deal with is the plug gap and the form of the electrode. A very important issue to appreciate here is that ignition performance is a function of spark

temperature and energy input. If the plug gap is very small, it takes minimal voltage to fire it and the energy involved is consequently low. From this, it follows that a bigger gap discharges a spark at a higher energy level, and that is exactly so. But a big gap uses coil output and, to an equal extent, causes a drop in the RPM to which the system can run before misfiring.

8-2. The longer the side electrode, the easier it can overheat. Also sparks propagate from edges easier than flat surfaces, hence the form of the modified shape.

8-1. The spark plug should run at a temperature that ensures deposit burn-off, and no more. Here are the factors that affect it.

A general rule here is that you should run as large a plug gap as possible, consistent with the engine’s peak RPM. Another factor here is that the bigger the plug gap, the more difficult it becomes for the plug cables to do their job before breaking down and firing sparks at random due to the higher resistance that the wide plug gap presents. Big plug gaps then require plug cables with a higher/better insulation value. Plug Tip The next issue to address, as far as plugs are concerned, is the tip form. Over the years, I have experimented with various designs, some of which have shown real promise. However, there is not space here to go into some of the complex designs and what they present in terms of performance advantages. You would have to make the tips in your own shop, and that is beyond the scope of this book. Instead, let’s look at a simple form that is about 99-percent effective compared to the best I have managed to date. Illustration 8-2 shows the electrode is both rounded off and shortened so it is only about a third of the way over the center electrode. This and the biggest gap possible (with good plug wires this can be as much as 0.055 inch) gets the job done performance-wise but erodes faster than a more conventional plug tip. This means changing plugs more often to maintain the spark quality. For what it’s worth, I install new plugs in a racer about every event.

Here is an Autolite plug I prepped for one of my engines. Follow this form, and you won’t go wrong. Plug Cables I so often see ignition problems that are fixable just by replacing the plug cables. Depending on the quality of the plug cable, the insulation

breaks down with time and heat. If the cables don’t look good, the chances are they are not. Cracked or burned insulation just does not work. Assuming they are not cracked or burned, the best way to determine if your cables are up to scratch is to check their resistance. For this, you need just a cheap multimeter. The commonly held belief—the lower the plug wire resistance, the better—is not quite true. Although counter-intuitive to the obvious, the cable needs some resistance. If, say, a plain copper wire is used with virtually zero resistance, it causes the spark to extinguish quicker, thus producing a less-effective spark. It seems that the best plug cable resistance lies between 100 and 1,000 ohms, so the margin here is wide enough to easily achieve.

Just because plug leads look new, does not mean they are functioning okay. Burning at the plug boots can be a big problem.

It is not necessary to spend a fortune on plug cables. On most of my budget-constrained engine builds, I use these Accel carbon-string plug cables. They don’t cost much, work fine, and are good for about two years of street/strip use.

A decent multimeter such as this can be had from your local parts store for as little as $20. It establishes the resistances of the plug cables.

These results show what a simple thing like old plug cables can do to output. The black curves are for the old plug cables and the red are for the new ones. Here we are looking at an average of 6 ft-lbs of torque and, at the top end, an almost 7-hp increase just by swapping old plug cables for new. The same applies to other items in the ignition system, such as spark plugs, rotor caps, and rotors.

If you have the budget, this spiral-wound wire-type plug cable is what you should be using.

The distributor cap discharge points and the rotor are subject to spark erosion. Install new parts when any significant erosion is seen.

In real life, plug cables can have a resistance far higher than we are talking about here. Typical carbon-string cables can be as much as 3,000 ohms per foot. While this is not exactly what we want, such a cable does do an effective job of eradicating RF radiation that otherwise interferes with your radio. Unless you are pushing for the ultimate no-matter-what, do not fixate on having the lowest-resistance wires there are. Just put in the best you can afford and locate the plug caps as far from (or protected from) a cable-killing heat source as possible. If you can afford better plug cables, then there is often a small output advantage to using spiral-wound-core plug leads. Tests I ran some years ago on a hopped-up distributor machine indicated that dropping the total resistance in the entire length of the plug cable below about 1,500 ohms showed no measurable increase in the RPM to misfire. Looking at the industry’s high-performance offerings indicates they mostly fall between the 45 to 300 ohms per foot range. This means any high-quality plug cable set meets the 100- to 1,500-ohm limits that are required.

This ZO6 Corvette engine has one coil per cylinder. Although appearing well equipped in terms of coil capacity, the system is designed to work optimally within the stock-RPM range.

Spark Erosion Next thing on the list here is: If your engine has a distributor, then make sure the cap and rotor are up to scratch. Spark erosion here can bring about a significant drop in spark quality. Also while you are checking/replacing the rotor, check the distributor shaft for side play. Anything detectable warrants fixing—shaft wobble causes ignition timing scatter, and on many engine types this causes a very real drop in output. The bottom line here is that the first step toward improving the spark capability of your performance machine’s ignition is to make sure all the consumable components are up to scratch.

With a 13.5:1 compression and 8,000-rpm capability, this Chevy LS6 engine called for a serious upgrade in ignition capability. If the stock components are operating properly but are not providing strong enough performance, you need to consider the aftermarket ignition options for improvements due to the needs of, say, the higher RPM, bigger plug gaps, higher compression, boost, or whatever. So let’s take a look at what you can do here.

Coils I am starting our system hop-up with coils, but let me first say that ultimately a coil’s performance is related not only to its ampere-turns ratio, but also to the manner in which the coil’s voltage is chopped at the moment

the spark is required. That said, we have to start somewhere, so let’s make it the coil and its performance. First, no matter how fancy the setup may look, factory coils are designed to provide sufficient spark for a stock engine, and many factory coils are not immune from deficiencies when it comes to increased RPM. So why do the OE coils not stack up here? Simple: The more output a coil has, the more current it uses, and the more expensive it is likely to be. Neither factor endears itself to an auto manufacturer trying to keep overall costs to a minimum. Aftermarket Coils So while factory coils may be more than acceptable for a stock situation, those high-performance mods you are contemplating can cause your engine to outpace the stock coil’s performance.

When used in an HEI distributor, this MSD coil was good for a 500-rpm increase before spark dropping occurred. Three factors enhance a coils capability: a greater turns ratio between the primary and secondary winding; a superior iron-alloy core; and a faster means of chopping and restarting the voltage/current fed to it. The first two factors involve coil design. By manipulating these, a coil with a greater capability can be produced, but in most cases you can expect the coil to use more current and/or cost more. Most aftermarket coils from reputable suppliers are worth about 500 to 1,000 rpm more than stock coils. The third factor—how we switch the coil on and off—matters greatly as far as RPM is concerned. We need to charge the coil’s magnetic field as fast and as long as possible between firing, to maximize stored energy. When the spark is required, we need to cut the voltage to the coil as rapidly as possible to maximize the rate of collapse of the magnetic field.

All this used to be done by the contact breaker points but, with the advent of electronic ignitions, it is all done with transistors and the like. Essentially, the electronics supplies voltage to the coil’s primary winding and then chops it at the appropriate moment. The key question here is: Can the electronics (an ignition box or the electronics associated with the stock computer) supply the requisite voltage up to and past the RPM capability of the engine? If there is doubt, you need to start checking to see which automotive electronics company has the required components to upgrade your system.

Ignition Timing and Curves Having a strong spark is only half the battle. We also we need to be sure the ignition timing is optimally set.

My own tests show that MSD’s high-performance electronic module and coil kit for the GM HEI distributor has an 11,000-rpm capability! So what is the optimal ignition timing? It varies and primarily depends on RPM and manifold vacuum, which, in turn, varies depending on the throttle opening. There are other factors involved such as fuel type, temperature, compression ratio, etc., but let’s focus on RPM and vacuum. At low RPM, there is significantly less mixture motion than required to speed up the progression of the burn after it has been initiated by the spark. Indeed, at low RPM, the burn of the fuel mixture progresses relatively slowly. But because the engine is turning slowly, there is more time for the burn to progress. To get the best output from a cylinder’s charge, the

combustion pressure needs to peak somewhere close to 15 degrees after TDC on the power stroke. At, say, 1,000 rpm, the best torque may be seen with the timing at, say, 14-degrees BTDC. But at 2,000 rpm, the timing need not start proportionally earlier. More advance will be needed but, at 2,000 rpm, the influence of mixture motion is starting to come into play. This has the effect of speeding up the burn, so something less than a proportional increase in timing is needed.

The RPM-induced advance, often referred to as mechanical advance, is tailored to suit the engine’s requirements by spring selection (yellow arrows) and the shape and mass of weights (white arrows). By the time a typical two-valve engine reaches 3,500 to 4,000 rpm, the increased charge motion/ turbulence is speeding combustion rates in proportion to the subsequently increasing RPM so no further advance is needed. This means that our ignition system has to advance the timing with respect to RPM relatively rapidly at first, but progressively reduce the rate as it approaches 3,500 to 4,000 rpm. This is referred to as the mechanical curve, if a conventional distributor is involved, because the advance is achieved mechanically by means of weights and springs. For an allelectronic system, we refer to this as RPM advance. Manifold vacuum also plays a strong part in dictating the optimal timing. The higher the intake manifold vacuum, the less the cylinder is filled. This in turn means less compression pressure is seen in the cylinders prior to the plug firing. The result is that the flame front progresses slower, thus requiring more advance. At part-throttle cruise, where the manifold vacuum is high, and the RPM still relatively low (say 2,000 to 3,000 rpm), a

typical engine may need as much as 50 degrees of advance to get the best results, and “results” in this instance means mileage.

Shown here are the basics of what you might expect of the RPM-related advance curves of three differently spec’d engines. The blue curve is representative of a stock engine with a modest cam and CR. The black curve is for a typical hot street engine; the red curve is for a race engine.

Modified Motor Timing Requirements When an engine is modified to any real extent, the timing curves of the RPM-related advance and the vacuum advance both need to be addressed. Hot, dense charges burn faster than cool, less-dense ones. A simple rule here is: When cylinder compression pressures go up, the advance required at that particular RPM comes down. Big cams reduce cylinder cranking and low-RPM pressures, so the RPM-related advance needs to start at a higher level and come on faster as RPM goes up. As the RPM reaches about 4,000

or so (depending on how big the cam is), the ramming effect brings about increased cylinder filling, so advance in excess of what a shorter cam might use decreases until total advance is about the same. If the CR is increased, cylinder pressures go up, so less initial and total advance is called for. In practice, increased cam duration and increased compression tend, to a certain extent, to cancel each other out. As for vacuum advance, this is all too often deleted from modified highperformance street and race engines. Making such a move does nothing to help total power output, but it does serve to increase part-throttle fuel consumption and reduce drivability. When I am modifying an engine, I go to great lengths to retain a functional vacuum-advance system. The return in highway fuel efficiency numbers alone is worth the effort. A big-cammed motor at, say, a 2,500 rpm highway cruising speed may need up to 50 degrees of total advance. Given the right advance under these conditions, the fuel mileage can be as much as 30-percent better. Throttle response is also much sharper.

Modified Ignition Systems In the 1970s and 1980s, there were more performance ignition companies pedaling their wares than spines on a porcupine. Time and product performance have weeded out most of the companies not worth dealing with. Over the years, I have dealt with a lot of ignition companies and the stories I could tell on that subject would fill a book. Suffice it to say that I have simplified my life by dealing with a few companies that have the stuff that works. MSD products are one of the best places to start. I have used MSD ignition systems since about 1972, so I have quite a backlog of experience. As a point of interest, I must be one of the few, if not the only, driver who has lost a major race due to MSD producing an even better output to the plugs than I expected. But that’s a story for another time. Suffice it to say MSD knows how to light-off cylinders very effectively. Another company I use very regularly is Performance Distributors. This company makes various high-output and high-RPM versions of the GM HEI distributor for use in GM and Ford V-8 engines (and a few others). When you buy a distributor from this company, it comes with custom curves (RPM and vacuum) to suit your motor spec. The price is good, the

reliability excellent, and the performance great. Best yet, it all comes in a self-contained unit with a one-wire hookup! Another company that has some promising ignition equipment is Ignition Solutions. They have a plasma generating unit, which turns any system’s spark into a multi-fire plasma discharge. The temperature, current, and energy put into the spark is dramatically increased. It even makes a multi-spark system, such as an MSD, fire multiple sparks. No, that’s not a misprint. Each of the MSD’s individual sparks becomes three or four sparks. So if the MSD sparked, say, four times before adding this unit, it now sparks a total of 16 times. I know someone is asking, “Where does the additional energy come from to generate these multiple and upgraded sparks?” So here’s how it happens. Plasma Booster Function In essence, the plasma booster could not be simpler, but it took the genius of Ulf Arens, a highly qualified university lecturer and ignition consultant, to come up with it. Installing the box, which has no direct connection to the vehicle’s power supply, creates a capacitive/inductive resonant circuit. The box has one ground wire and one wire for each coil positive terminal. For a single coil, the hookup involves connecting just two wires. In the case of the multi-coil V-8, there is a simple nine-wire hookup. When the ignition system fires the plug, the capacitive/inductive part of the system resonates. This first fires the plug in the normal direction in a conventional manner but, a fraction of a micro-second later, a spark in the reverse direction also occurs. This swinging of spark polarity delivers four to six high-energy sparks before the available energy drops to a level too low to be of further use. The considerable increase in plug current and energy has to come from somewhere. Fortunately, the source of this increased spark current/energy is easy to appreciate. Take a look at the result of an output of a conventional ignition system in Illustration 8-3. The 1- to 2-microsecond spike indicated by arrow A is the high-temperature (about 3,300 degrees F) part of the spark that actually lights the charge. The section indicated by arrow B is the after-burn phase. It is of relatively low temperature and contributes nothing to the ignition process. The amount of energy in area B of the spark is actually about 85 to 90 percent of the total spark energy and, in a conventional system, just goes to waste.

Here is a Performance Distributors HEI-style unit in a race-winning bigblock Chevy built in my Charlotte, North Carolina, shop. Note the clear cap, which makes it easy to check the rotor alignment. After alignment is verified, we usually change to a regular solid-color cap.

The Ignition Solutions plasma-generating box has LEDs (arrows) that indicate each cylinder is receiving the plasma output.

8-3. The spike (arrow A) is the coil output into a conventional spark. Section B in a conventional system just goes to waste. When a plasma booster is installed, the otherwise-wasted energy in area B is utilized in the multiple spark the unit delivers. Not only are the sparks multiple, but each spark delivered also has about 80-percent-more current than a typical spark. All this has obvious advantages in terms of power potential, but plug life and ignition reliability also come in for some improvement. Since the plug’s spark polarity switches back and forth, the erosion of the electrodes is almost canceled out, so the plugs last much longer. Also in the very unlikely event of the plasma booster failing (it is about as reliable as a 6-inch-steel rule), the system automatically reverts to the OE ignition without any human (or otherwise) intervention. N2O and Boosted Engine Timing Any time the density of oxygen is increased, the charge burns faster. As previously noted, peak power is usually produced when the cylinder pressure peaks at about 15 degrees after TDC. If the charge burns faster, then there is a real need to retard the point at which the charge is fired—by pulling out some timing. As a basic rule, if you expect to increase output by more than one-third of the engine’s original output, either by means of nitrous or some sort of supercharging, then expect to back out the timing so the total timing at WOT is about 5 degrees less. As boost or nitrous is

increased, expect this trend to follow suit. A heavily boosted or nitrousinjected engine may need as little as 20 degrees of total timing.

Expected Results The key ingredients for a super spark are temperature and watts of power. High currents and voltages generate this spark intensity, so that is what is needed when you buy a performance ignition system. But all the effort of accruing the parts for a killer ignition needs to produce two things: a greater output and lower fuel consumption. In terms of spark output on a test rig, I have seen very strong results with an MSD box and a plasma booster tied-in to work together. The question is: How much extra in terms of power and economy did it produce? Well, let’s start with the old adage that you cannot fix something that is not broken. Some engines have a very good burn characteristic that needs only (relatively speaking) a minimal spark to get the job done. Others have chambers that do not have such good characteristics. If you put a maxed-out system on an engine that lights-off easily, the gains will be minimal, even though the ignition system is working exactly as it should.

Here is what a plasma booster can do for a spark. The lower example is that generated by a high-output competition ignition system. The upper example is the result of adding a plasma booster to the system that produced the lower spark.

A prime example here is the original Mini Cooper engine (A Series). If this engine has a decent spark from an accurately firing but basic electronicignition system, we find that adding bolts of lightning does nothing to the output or economy. But engines like this are not the norm. In practice, we find that four-valve, pent-roof engines usually respond to a super spark to the tune of 2 to 3 percent. That said, I did find a 5-percent increase on a four-valve Ford 4.6 Mod Motor—sometimes those bigger gains are there. Engines with big bores (big blocks of almost any make) also usually respond well because a more-aggressive light-off leads to a faster initial flame-front travel, and that’s a good thing. Finally, some combustion chamber shapes do not burn well when a high piston dome (for compression) figures into the equation. Big-block Chevys and Chrysler Hemis fall into this category and usually respond well to a big, high-intensity spark.

CHAPTER 9

REAL CAMSHAFT SCIENCE Selecting a cam or, more accurately, deciding the cam events that the engine needs has traditionally been regarded as a black art. Most publications feed you just enough info for you to get into trouble (in most instances) or for you to decide you should let a tech guy at a cam company make the decision for you. Either way, this can be bad. Writers don’t, in 99 percent of the cases, have the knowledge to educate you to a point where you have a working idea of what’s needed and why. In the case of the cam company’s tech guy, you are at least dealing with a person who has, in most cases, a lot of experience with what works, but not necessarily why it works. If you want proof of this, you can rerun a little experiment that I did with some friends. We took an engine that had already been built and dyno’d to determine the optimum cam. At this point, four major cam companies were anonymously called for a cam spec for this engine. There were two stipulations: The cam had to be a flat hydraulic, and the intake duration (at 0.006 follower lift) had to be 280 degrees. We received four very different cam recommendations. Among these was an overlap variation ranging from 52 to 64 degrees, and lobe centerline angles (LCA) ranging from 110 to 114. Also, single- and dual-pattern cams were both put forth.

The cam and valvetrain components chosen are critical to engine’s success as a power producer. The subject of optimal valve events has been poorly covered in countless books and magazine articles. In this chapter I will start you down the road toward what will actually work first time around in your engine. This prompts the question: Can they all be right? Certainly not, but the situation actually took a turn for the worse when we called three months later for a cam spec for the same engine. In three out of four cases, we got a different tech guy but in all cases we were given a different cam spec than the original. So we now had eight different cams, all supposedly near optimal for the job. The real kicker here was that dyno testing had shown that, for our engine (as opposed to a generically similar one), a single-pattern cam on a 108-degree LCA produced the best power curve (highest average output over the used-RPM range), yet not one of the cam companies had recommended a 108 LCA cam! To be sure, all the suggested cams would have produced far more output than the stock cam each would have replaced. However, none would have produced the near-optimal results of the cam finalized by the dyno testing. The point I am coming to here is that in spite of the fact the four-cycle internal combustion engine has been with us for more than 100 years, there is still a great shortage, even within the cam industry, of knowledge about what valve events work for a given engine spec. (In the next three chapters, which all pertain to the valvetrain, I will supply the information you need to

know to make a top professional judgment when it comes to spec’ing-out a cam for your engine.) So what qualifies me to be this critical of previously printed matter and the cam industry in general? Simple: It’s better than half a lifetime’s worth of dyno testing cams (more than 45 years) to draw on and a pile of race and championship wins, pole positions, fastest laps, etc. And, on top of that, achieving sufficient expertise on the subject to teach it at university to some of the top winning-race/engine-building pros in the business. With all that said, let us now get down to the promised “real” camshaft science.

Mechanical Attributes Simplified Let’s be clear about things here. Although I can simplify the basics of cam/valvetrain dynamics, what you learn here won’t come close to getting your university degree in cam dynamics. The subject is unbelievably complex. What I do is give you the basics, which will open the door to a far greater understanding than you probably have now.

9-1. Here is what is seen looking at a pushrod V-8 cam end on: intake lobe lift (1), exhaust lobe lift (2), intake duration (3), exhaust duration (4),

overlap (5), lobe centerline angle (LCA) (6), cam advance (A) and retard (R) (7). For a multi-cam engine, the advance/retard and LCA is phased by the valvetrain gears. First, let’s look at the camshaft’s basic descriptive attributes. These are shown in Illustration 9-1 and represent the starting point of understanding how to time a cam in, why certain factors change its characteristics, and how optimally it may work. The most important factor to understand here is that the lobe centerline angle (LCA) and the duration of both the intake and exhaust directly affect the overlap produced.

9-2. Shown here is an intake cycle from start to finish. The duration (in this example, 270 degrees) is indicated by arrow No. 3. The degrees

opening before TDC is indicated by arrow No. 1 and the degrees after BDC to the closing point by arrow No. 2. The No. 4 line indicates the intake centerline angle, which is exactly halfway between the opening and closing points. Illustration 9-2 shows the points where the intake valve opens and the position of the piston in the bore with a cam of 270-degrees duration at the lash point. The blue partial circle indicates this duration. Illustration 9-3 shows just how you see it on a cam spec card. The next move is to take a look at the valve lift (or, as cam designers call it, the lifter displacement curve), produced by the cams form. This is shown in Chart 9-4 (note the caption comments on the duration at various tappet lifts). If you have absorbed all that, what we are going to do now is take a lift curve and dissect its dynamics. In designing a lift curve, a cam designer has to consider not just the lift imparted to the tappet (or lifter or follower) but also the velocity, acceleration, jerk, and jerk II. If you are Harvey Crane, Jr., then this list goes on to include snap, crackle, and pop! All tappet motion is quoted in terms of crank degrees, so lift is “so many thousandths of an inch” at “so many degrees” from the starting point (i.e., 0.045 lift at 30 degrees from the starting point). Velocity is not quoted as a particular speed in inches per second (because engine speed changes) but rather in thousandths of an inch per degree of crank rotation. The velocity is the “rate of change” of position or displacement of the tappet in relation to the motion of the crank in degrees.

9-3. By melding the intake- and exhaust-duration arcs (at the top) together, we form the valve-opening event (at the bottom). In this example, the inlet opens (IO) 25 degrees before TDC and closes 55 degrees after BDC. The exhaust opens (EO) 55 degrees before BDC and closes 25 degrees after TDC. The way we quote the events of this cam is: 25-55-5525. The overlap, when both the intake and exhaust are open, is indicated here and, in this case, is 50 degrees (25 + 25). It is important to understand this rate-of-change factor, so we can move toward understanding the remaining motion factors mentioned earlier. The “tappet acceleration” is the rate of change of velocity, and “jerk” is the rate of change of acceleration, and so on. It is important that each of these factors changes progressively; otherwise, the valvetrain system suffers spurious vibrations and at some point goes into resonance. When that happens, the valve motion seriously deviates from that intended by the cam form. For

instance, if the acceleration instantly went from zero to some relatively high value, it simulates the valvetrain being hit with a hammer. Such a move can hardly be expected to produce a quiet or otherwise functional valvetrain. There is a simple experiment to demonstrate how this works, and you can do it on your next drive. When you want to bring the car to a stop, put on the brakes by applying, say, 40 pounds (or whatever) of force to the brake pedal. Do not change the applied force until after the vehicle has stopped. What you are doing here is slowing the car at a constant deceleration. When the car stops, there is an instant change of acceleration from some value (say, 0.2 g) to zero. Notice when the car stops, you are jerked by that instant change in acceleration. When we drive, we avoid that jerk at the stopping point by letting off the brake pedal progressively as the vehicle comes to a stop. If you practice the art of smooth driving, you would be progressively applying the brake so, at both ends of the stopping procedure, the change in deceleration between zero and peak is a smooth progression. If you understand the dynamics of this example, you are well on the way to understanding the interrelations of displacement, velocity, acceleration, jerk, etc.

9-4. Duration No. 1 is at a solid lifter’s lash point. The lash point at the lifter is the lash at the rocker divided by the rocker ratio. Arrow No. 2 is the so-called advertised duration and is usually 0.006 inch for hydraulics and 0.020 inch for solids. The red arrow indicates the duration at 0.050 inch. The whole point of refining the cam’s profile in this manner is to make its operation smooth. Doing so considerably reduces the propensity of aggravating the valvetrain as a whole or any of its components at their primary resonant frequencies. Some of the advantages are the need for less spring to control the valvetrain to its design RPM, less valve-to-piston clearance without hitting the piston, less chance of spring surge below peak RPM causing valve bounce, and so on. We should add valvetrain longevity to that list as well.

Hi-Perf Four-Cycle Engine The type of engine we are dealing with here is traditionally known as a four-cycle engine. While that may be largely true for many daily-usage production engines, with exhaust systems of less-than-adequate flow and scavenging capability, it is certainly not true for a genuine high-performance engine. Given that a length-tuned exhaust system can produce a very strong negative-pressure pulse at the exhaust valve, we see that there is an opportunity to produce an induction phase wholly independent of the piston’s motion down the bore.

Here is what the motion data of a typical cam looks like. The red curve is the lift from the base circle on up. The black curve is the lifter velocity (rate of change of lifter displacement) in inches per degree. The blue curve is the acceleration (rate of change of velocity) in inches per degree per degree (usually seen as inches per degree2). The green curve is the jerk (rate of change of acceleration, or inches per degree3).

9-5. Starting at cycle No. 1, the exhaust-generated vacuum starts the intake charge moving into the cylinder way before the piston even starts down the bore. As the crank rotates farther, we get to cycle No. 2. This is normally considered the charge-inducing stroke. In an ideal situation, cycle No. 1 has cleared the combustion chamber and put a considerable amount of kinetic energy into the incoming charge before the piston starts down the bore. The result is an engine that can achieve a volumetric efficiency well over 100 percent. The bottom line is: A good exhaust system is worth a lot of extra torque, horsepower, and, best of all, extra mileage. But to make all this work as intended, the cam must generate the right events around TDC. For a well-tuned system, the exhaust pulse can be super effective over as much as 4,000 rpm. In terms of intensity, it can be far stronger than the suction caused by the piston traveling down the bore. A really good twovalve race engine making upward of 135 hp per liter can have as much as 150 inches of water suction on the intake valve at TDC, while the piston, at

peak power, has only about 20 inches. Any more than that and it is a strong indicator that more cylinder head airflow is needed. Having returned to the subject of airflow, let’s focus on that and investigate how cylinder head airflow and port velocity should be managed for best results.

Airflow Dynamics After years of studying the subject, I realize that one basic rule emerges as the most important to appreciate if optimal output is to be seen from any high-performance engine. The question we need to ask ourselves here is: which of the gas-moving cycles is the most influential in terms of generating output? Referring to our five cycles in Illustration 9-5, we have numbers 1, 2, and 5 to consider. While none of them is without its own degree of importance, we do need to establish a priority. Up until maybe as late as the 1980s, the delayed intake closing after BDC was considered the biggest influence in making power. Although this was a widely held belief, the reality is that it is the overlap period that is the number-one influence on the cam’s success. This does not mean that more is better. What it does mean is that it is very important to have the right amount of overlap for the RPM involved and the right proportion of the intake and exhaust duration occurring within that overlap. Getting the overlap right for the engine geometry/flow characteristics concerned means first taking optimal advantage of exhaust-driven combustion-chamber scavenging, and then having the intake valve as far open as possible at the start of the pistondriven intake stroke.

This 302 Ford 5.0 Mustang makes 475 streetable hp. This was mostly due to a strong interaction between cam events and exhaust. Cylinder pressure tests and subsequent dyno testing prove that the suction caused by the exhaust, and what the piston does on the intake port in the first half of the induction cycle, is critical. If this is not done right, there is no chance of rectifying the situation in the second half of the stroke. Trying to fix an inadequately filled cylinder in the second half of the induction cycle is a bit like bolting the stable door after the horse has left. Having made that point, it should be evident that opening duration, which is the most-often-used cam selection parameter, is actually the wrong way to go. Since the huge depression caused by the exhaust brings about the single strongest action on the intake port, it follows that the overlap to allow this to happen must be the most important element of all the opening and closing events to get right. This being the case, it demotes duration as a deciding factor to several places down the list.

Be aware that there is much more to cam selection than just plugging in a bigger cam. When buying a cam, consider that one with the right events because it costs no more than one with the wrong events.

Tuned-length intake systems look high-tech, but at the end of the day it’s the exhaust working with appropriate cam events that makes for results.

The more accurate the cam is spec’d-out, the more important it becomes to time it in right. Chapter 10 explains why. Based on this, we can then say that the number-one selection factor is overlap. Following this, the number-two selection parameter is the lobe centerline angle. On a single cam engine, the LCA has to be ground into the cam. But for an engine with separate cams for the intake and exhaust, the LCA is adjustable by means of the timing on each of the cams in relation to the crank. It could be said that duration is the third most important factor but, in reality, we cannot independently adjust it, so it’s really only a by-the-way deal. Here’s why: After the overlap and LCA requirement (which we deal with soon) has been determined, we find the duration is now fixed. In other words, × overlap with Y LCA can only produce Z duration. Assuming a single-pattern cam, the duration is the sum of half the overlap, plus the LCA, multiplied by 2. An example looks like this: Let’s say the overlap required is 54 degrees and the LCA is 108 degrees. Half of 54 is 27. The 27 plus 108 is 135, which, when multiplied by two, comes to 270. That duration is the only one that gives the overlap and LCA called for. From this, you can see that the duration is totally dependent on the overlap and the LCA being used.

Choosing a Cam Up to this point, you probably were buying into a cam selection procedure based on duration. Even if you did not know exactly what you were doing, you could at least have made a decent ballpark guess if duration was indeed the sole deciding factor. As of now you have probably learned that what you thought you knew was in fact relatively useless, and the procedure to select a cam is based on variables you might not have previously considered. This means if you had little doubts about your ability to select a cam before you started this chapter, you probably really have them now. But in the next chapter, we fix that.

Having the correct events is just the start of the cam selection procedure. At the end of the day, the entire valvetrain needs to be spec’d-out.

Shown here is the real-life lobe centerline angle for a small-block Chevrolet roller-follower cam.

Assembling the components for an ultra-functional valvetrain is all a question of attention to detail. A point to note here is that any valvetrain is only as good as the valvespring. Assuming a smooth cam profile, the rest of the job is very much making sure that spurious vibrations do not get the better of things; this is where aluminum components work well.

As unlikely as it may at first seem, the crank damper is actually part of the valvetrain and can strongly influence its dynamics.

CHAPTER 10

CAM EVENT CRITERIA We are still on the same subject here, as in Chapter 9. We are going into more detail on valve-event timing and how the spec of the rest of the engine affects optimal events. The topics we need to cover, in order of importance, are: overlap, lobe centerline angle, duration, and valve lift. What I want to do here is to help you avoid buying a generic cam. That’s one where the cam catalog tells you it’s a good cam in a Chevy small-block with a displacement between 265 and 440 cubes. If you had cause to believe that one cam could be optimal for such a wide range of displacements, then you are going to find this chapter a real eye-opener.

Overlap Determining how much overlap is required for the job at hand sets the scene, so to speak, for the rest of the cam spec selection. Just how much should be used depends on what the engine’s primary function is. If it is truly a street engine, then idle quality, low-speed torque, and cruise fuel consumption are almost certainly high priorities. On the other hand, if it is a race motor, outright performance takes precedence over everything else. At the street end of the scale, small overlaps are called for, while the race end needs large amounts of overlap. This means the first factor we need to take into account is the primary function of the engine involved.

The mechanical dynamics of the valvetrain is the focus of the last chapter. Here we will start on the gas dynamics needed to achieve maximum output. For a production engine intended to power a daily driver, overlap values of 10 to 35 degrees are common. Such amounts of overlap favor idle, mileage, and emission considerations. If we consider typical-performance street cams for an engine that’s most likely equipped with an efficient intake and exhaust, we find that overlap values that still satisfy true street requirements can typically fall between 30 and 55 degrees. At the other end of the scale, all-out race cams can have anywhere between 85 and 115 degrees of overlap. So far, all I have done is give you a vague clue as to what your engine might specifically need for overlap. That’s no better than you could have obtained from many other sources. Obviously, I need to do better than that, so let us look at some key overlap factors in an effort to more accurately determine what might be needed. In one sentence, here is a summation of what is a controlling factor: The characteristics produced depend on the overlap triangle area and intakevalve flow in relation to the capacity of the cylinder it feeds. Because you

really need to have equipment to measure cam profiles very accurately (such as Audie Technology’s Cam Pro) and a flow bench, it is necessary to simplify this. But even with this simplification, you are way ahead of where you might have been. In simple terms: The overlap required is a function of the intake-valve diameter and lift, and the overlap period in relation to the cubes the intake has to feed. For a given cylinder head, the bigger the displacement of the engine, the more overlap it can have for given low-speed and idle qualities. For instance, a 350-ci V-8 idles perfectly with 40 degrees of overlap. Stretch that engine to 406 inches by means of boring and stroking, and the same low-speed attributes can be had with about 1 degree more overlap for every 5- to 7-ci increase in engine size. This is a good general rule for most twovalve V-8 engines.

The amount of overlap used for this 740-hp 355-inch Dodge is dictated by the need for maximum horsepower. Its idle quality and low-speed manners are far from what is needed for a street motor, as is the expected valvetrain life.

This Audie Technology unit measures cam-profile dynamics and determines dynamic quality.

When it comes to camshafts, quality of manufacture is especially important. This Lunati cam not only checked-out on my profilemeasuring equipment as professional race quality but also the rest of the cam was of premium-grade manufacture.

On the other side of the coin, we have a situation where the engine’s capacity has remained unchanged, but a big valve head(s) has been installed. This situation calls for a reduction in overlap, if the same idle quality and low-speed output are to be preserved. In the range we are looking at, we can say that for an increase in the intake valve of, say, 5percent, a reduction in overlap area of a like amount is about what is needed. As a starting point, use the values shown in Chart 10-1. Some examples here can help you make a sound choice, which keeps subsequent determinations on the mark. Let’s assume we are spec’ing-out a performance street cam for three different engines: first, a 2-liter, four-valve-per-cylinder engine; second, a 350-inch two-valve small-block Chevy; and last, a 427-inch small-block Chevy (i.e., a big-bore-and-stroke version of the 350). Section # 2 on the chart is the range we are referring to here. Because the four-valve engine has a lot of intake-valve diameter (and consequently circumference) per cubic inch of cylinder displacement, it presents a lot of breathing area to the cylinder at low lift and opens up to high flow very quickly. A typical four-valve engine has only about 11 ci of cylinder to fill for each inch of valve diameter. As a consequence, the overlap needs to be on the short end of this range, i.e., about 35 degrees.

10-1. This chart shows a typical spread of overlap for various applications. The middle of each range represents the amount for a typical two-valve engine having 21.6 ci of cylinder displacement for each inch diameter of intake valve. This puts multi-valve engines at the lower overlap of each range and big-displacement under-valved engines at the bigger overlap values of each range. Section No. 1 is for engines requiring best low-speed output and mileage. No. 2 still caters for low speed but does give away a little for additional higher-speed output. No. 3 represents a common range for high-performance two-valve street engines. No. 4 is the range for most endurance race cams. No. 5 is the amount of overlap generally reserved for big-displacement two-valve engines that generally have too many cubes for the size of valves. For our typical 2.02-inch-intake-valved 350 Chevy, we have 21.6 cubes per inch of valve and this calls for a greater overlap than for the four-valve engine. Here, about the middle of the range works fine, so some 45 degrees should be used. Now let’s take the 427-inch example with the same heads as the 350 but just more inches. In this case, each inch of intake valve has to feed some 26.5 cubes, so the amount of overlap needs to be increased. Indeed, to meet the same requirements, this intrinsically under-valved engine needs about 10 degrees more overlap than the 350, thus putting it at the big end of the range of section # 2 on the chart. A word of caution here concerning your choice of overlap: Do not fall into the trap of thinking if you can successfully use a certain number of degrees, then adding a few more won’t really cause any real low-speed problem. In reality, that is not quite so. The reason is that when the overlap angle is increased, so is the amount of lift seen at the valves at TDC. What you are doing is increasing the area. A 10-percent increase in overlap angle is about a 21-percent increase in actual overlap area, and that is what the engine sees. This being the case, unless you know better from experience, it is better to err on the conservative side.

Because of its significantly bigger 2.23-inch intake valve, instead of the more usual 2.02-inch intake, this small-block Ford head requires about 10 percent less overlap to produce the same results.

Big-bore blocks (such as this unit from Dart) can, in conjunction with a stroker crank, boost the cubes of a small-block Ford Windsor, from 351 all the way to 460. Without a corresponding change in valve size, this calls for much more overlap to be used. So much for how the overlap may change with changes in engine and valve sizes. Your move now is to make an assessment of the overlap your

application is likely to need.

Optimal LCA The realm of cam LCAs is a largely misunderstood aspect of camshaft valve events. My intent here is to fix that, so you have a realistic understanding of what is going on. After you have absorbed what I have to say, you are likely to find that you are at odds with 99 percent of cam industry professionals. Why? Because what I am about to tell you is still regarded with suspicion by many in the cam industry. This is in spite of the outstanding success I and all those who have used it have had with this technology. Before going into details, let me make one thing clear: The LCA is not the adjustable feature that most cam companies would have you believe. They make comments like, “Well, a 108 LCA is too tight for the street because it kills vacuum and idle quality, so you need a 114 LCA for this application.” If you hear something along those lines, then you are, to some extent, being misinformed.

The A-series engine that powers the original Mini Cooper has siamesed intake ports (two ports for four cylinders). As overlap values increase, the propensity for the second cylinder to draw on any one port causes robbing of the leading cylinder on that port. As a consequence, each cylinder requires a different overlap for best results.

A simple vacuum gauge, such as this example from Autometer, provides a valuable means of setting up the idle and the cruise-ignition timing calibrations to get the best low speed and idle from an engine. There is an LCA that, in terms of output, delivers the best area under the curve—it is pretty much fixed, except for very short or very long duration cams. Sure, a degree or so one way or the other may not be out of line, but 4 degrees certainly is. Let me give you a starting point example so you can at least see the extent of what we are dealing with. A small-block Chevy with a set of heads having 2.02/1.6 on the intake/exhaust valves, with a 10:1 CR, requires a 108-degree LCA for the best torque/horsepower. This holds for a street truck engine or a bracket racer expected to get up on the cam at 3,000 rpm. That means any cam paired with a 10:1 CR, from about 245 degrees (at lash) to 310, works best with the 108 LCA. If you are told that the cam you need should have a wider LCA, it is because you are trying to run more duration than the application calls for! So what does affect the optimal LCA for a given engine? Although far from the only factor, it is basically the amount of breathing area presented to the cylinder in relation to the cubic inches of that cylinder. This, in itself, is a complex issue, but we need to start somewhere. It will serve you well to

see how the optimal LCA can change—with low-lift breathing capability tests I did back in the early 1990s. At the time, I was working with a company that had produced a pushrod four-valve head for a small-block Chevy. My task was to establish just what these heads could produce in terms of additional power. To this end, I tested with a set of reconditioned, early high-performance factory heads; a set of ported aftermarket aluminum heads; and a set of the four-valve heads, which I had also ported. The flow difference of each of the two-valve heads was minimal in the first 0.150 inch of lift. Above this, the ported two-valver pulled away in terms of flow by some 65 cfm at 0.700 lift. Both these sets of heads produced the best output on a 108 LCA. The four-valve head had, by comparison, huge low-lift flow and produced the best output with a cam on a 112- to 113-degree LCA. All cams used the same profiles, so were of the same duration. Just for the record, the four-valve heads topped the ported two-valve heads by some 180 hp!

For the best idle quality and economy from fuel calibrations, a good O2 mixture analyzer is vital. This is a unit from Autometer. This (and many other similar tests) brings us to the first conclusion: The greater the low-lift flow is in relation to the displacement of the cylinder being fed, the wider the LCA needs to be. The other side of the coin here is that the bigger the cylinder fed by a given head, the tighter the LCA required. This means that if you bore and stroke your engine, and if the LCA was optimal before, it now needs to be

tightened up. An example for this is that if a 350 small-block Chevy is increased to 383 inches, then, if the same heads are used, the optimal LCA goes from 108 to a tighter 106. So rule number one is: More inches or less low-lift flow requires a tightening of the LCA, and more low-lift flow or fewer inches requires a widening of the LCA. The next most influential LCA factor is the engine’s CR. Increasing the CR causes the optimal LCA to spread, and decreasing it brings about the reverse. I have been in several LCA debates. Some proponents of any number of untested theories put forward the fact that a 500-inch big-block Pro Stock engine produces best power on 112 to 114 LCA. They ask me why I am using 106 LCA on my hopped-up 480- to 511-inch street bigblocks. This is one of those elements of the LCA that is rarely appreciated. In fact, the higher the CR, the wider the LCA needs to be. If we drop the CR of a typical Pro Stock engine from 16:1 to about 10.5:1, it needs a significantly tighter LCA. Even with an intake valve on close to 2.5 inches, such a motor would need an LCA of about 108 to 109 at the widest. The next concept for understanding what the optimal LCA may be is the intake-valve acceleration. The faster the intake is accelerated off its seat, the wider the LCA needs to be. If your engine had an optimal LCA cam and utilized 1.5:1 rockers, then the use of, say, 1.7:1 rockers could mean that getting the best from the ratio increase requires spreading the LCA by as much as 1 degree. LCA: How Critical Is It? If you pick up a typical cam catalog and peruse the off-the-shelf cam specs, you might think that the LCA cannot be that fussy. Even for one engine description there are half a dozen different LCAs proposed. In Chart 10-2 is an LCA test performed on a small-block Chevy engine. Just for the record: I have done such tests on Mini Cooper engines right through to Pro Stock and even on a half of an Formula 1 engine (it’s cheaper to test on half of one than a whole one). I have done these tests with big cams, small cams, high compression, low compression, and most other combinations that look like they reveal what is really going on. Two factors have become evident: The first is that there is a narrow range of LCAs over which the engine delivers its best and, second, it is better to err on the tight side rather than the wide side.

This can be seen from the curves in the chart. With the 112 LCA, the torque curve was substantially down, and going to the 110 LCA raised output just about everywhere. The 108 LCA gave the best peak horsepower by a small margin and the second-best torque output, trailing the best (106) by only a small margin. The 108 LCA was, however, better at the lower RPM. The point here is that, on the track and over the RPM range used to race, there would be no measurable difference between the 106 and 108 LCA. There would, however, be an easily seen negative impact on performance if the 110 LCA had been adopted. Too tight an LCA on the engine produces overall negative characteristics, such as reduced idle quality and low-speed output. Just how much of the low-speed manners have been lost depends ultimately on the overlap used. For the test shown here, a hydraulic flat-tappet single-pattern cam having a moderate lift and a duration of 278 degrees was used. With very-high-compression engines, such as used in Pro Stock, my tests to date indicate that having the LCA up to 2 degrees too wide is not as detrimental as it is with lower-compression engines.

10-2. What you see here are the torque curves for LCAs of 112 (black), 110 (darker green), 108 (broad light green), 106 (purple), and 104 (red). All these tests were conducted in a small-block Chevy with 2.02/1.6-inch valves and approximately 9.5:1 CR. The tests were conducted at an acceleration rate of 600 rpm per second. Each cam had the fuel and timing optimized.

Chart 10-3 highlights the price of not having an optimal LCA for a given specification. The engine here is a small-block Ford 302 that becomes the recipient of a 3.4-inch-stroke Scat crank to replace the 3-inch-stroke stock crank. This bumped the already 0.030 overbored engine from 306 to 347 cubes. With the stock stroke and 306 inches, the engine produced its best with the cam on a 110-degree LCA. After being rebuilt with the 3.4-inch-stroke crank, the engine was retested using the same 110-LCA cam. As the curves show, the engine produced a sizable increase in output from the extra cubes. Using the same profiles, a new cam was ground on a 108 LCA. As can be seen, this delivered almost 20-hp and 20 ft-lbs improvement, with no downside in idle quality when compared to the original 302 on a 110 LCA.

10-3. This before-and-after stroke change test shows the value of choosing an optimal LCA for the engine spec concerned, and not just using some generic grind. Here, it is worthwhile to just take a look at why idle and low-speed cruise vacuum is influenced so much by cam duration for a V-8 engine, especially when a single-plane intake manifold is used. Illustration 10-4 explains the situation and should bring home the need to not overdo the use of overlap in the first place. Determining the LCA Earlier in this chapter I outlined a few basic rules that dictate how the optimal LCA changes with engine spec. Rule number one was the relationship between cylinder displacement and low-lift flow. Without the aid of a flow bench, and some math rather more complex than I want to show here, we find that it’s best to take a step back and use valve size to determine the LCA required. It is not quite as accurate, but it is at least an order of magnitude better than the guesstimates used by even the most experienced engine builders. To painlessly get a base LCA, refer to Chart 10-5 as a starting point. Using this chart, you can get an LCA figure for any parallel two-valve engine using a CR between 9.5:1 and 10.5:1. This gives us a base value as a starting point to build on. However, when engine specs do not exactly fit, the same spec changes are necessary. First, let’s look at the effect of the CR. For every ratio above 10.5:1, add 0.75 degree to the LCA. If we apply that to a 15:1 drag-race small-block 355-inch Chevy with an intake of 2.23 inches in diameter, we get 110 degrees for the base figure, plus 3.375 extra degrees for the higher CR. This works out to a 13.375 LCA. In practice, on a 900-hp drag-race 355, the best LCA was between 113 and 114 degrees. Up to this point then, we find that in terms of LCA prediction accuracy, things are looking good. Now let’s put valve angle into the equation.

10-4. The blue circle represents the flow area past the carb butterflies. With a single-plane manifold, all the cylinders are interconnected. When one cylinder is in the overlap period (represented by the left-hand cylinder), there is a flow path from the exhaust port right on through the combustion chamber, out through that cylinder’s intake port, and on into the intake manifold. When any particular cylinder is in the overlap period, one other cylinder is near maximum piston speed and in the middle of its induction stroke. If the cam has minimal overlap, the through-flow area is relatively small (represented by the small red circle). Under these circumstances, there is a minimal combined in/out flow area into the intake manifold. This produces a mixture with the minimum of exhaust dilution and a maximum vacuum reading. If the cam has a large amount of overlap, then the through-flow area from the overlap on any one cylinder back-flows into the intake manifold. This reduces intake vacuum, but the situation is further compounded by the need to open the carb butterflies farther to get a combustible mixture. The effect of low idle and cruise vacuum can be reduced somewhat by using a two-plane 180degree intake manifold; this separates the induction and overlap phases of the cylinders.

10-5. This line represents what is needed for a typical two-valve-percylinder, pushrod engine having a compression ratio between 9.5:1 and about 10.5:1. Using this as a starting point, we can make changes to the LCA brought about by other changing factors, such as CR, intake valve acceleration, and valve angled to the bore, etc. The first step in using this chart is to determine the cylinder volume versus the cubic inches involved. Here, we use a small-block Chevy as an example. Let’s take some small-block Chevy displacements and see how they work out in practice. The cylinder displacements of a 302, 350, and 383 are 37.75, 43.75, and 47.87 ci, respectively. Assuming the typical 2.02inch valve, this gives 18.69, 21.66, and 23.7 ci per inch of valve diameter. To establish the LCA for each of these displacements, go up the left-hand side of the chart to the value being used. Assuming the 350 here, we will go to the 21.66-cubes-per-inch value, then go across the chart to the line, and, at that point, drop to the bottom scale. This results in an LCA value of 107.5 degrees. Rounding this off to the nearest degree means either a 107 or 108 LCA. In practice, a 108 works very well. In the same way, we get numbers of 111 for a 302 and 105.5 for a 383. For 383 builds, I most often use a 2.08-diameter intake

valve, giving 23 cubes per inch. On the chart, this gives an LCA of 106 degrees. In practice, tests have shown a 10:1 383-ci engine with this size valve delivers its best output curve on a 106 LCA, so you can see the chart predicts (for a parallel two-valve engine) to an accuracy of about 1/2 degree. If the valves are inclined, as with a big-block Chevy or a typical fourvalve engine, then cross flow between the intake and exhaust can take place easier and upset both the idle and low-speed performance. When valves are angled, concessions have to be made. Also when the CR changes from the valves the chart was based on, changes have to be made. Our example is the ever-popular big-block Chevy. In practice, the dyno shows that the LCA needs to be 2 degrees wider than the chart predicts for parallel valves. Plugging in the popular size of 496 cubes with an intake of 2.35, we find the chart predicts an angle of 102 degrees. Add 2 degrees to compensate for the valve angles, and we have an answer of 104 degrees. This works well in practice for a 10.5:1 big-block Chevy, but valve-topiston interference can be an issue with bigger cams, and that may dictate a slightly wider LCA yet. However, a big-block Chevy loves compression and 12:1 produces much better results than 10.5:1. When corrected for the increased compression, this works out to a 106 LCA. In practice, I have found that 106 to 108 works very well. Taking the case of an inclined four-valve-per-cylinder engine, we find that the LCA can be determined in much the same fashion. First add the diameters of the two intake valves together and divide that into the cubes per cylinder. Using a popular 2.2-liter engine with 1.26-inch intake valves (12.88 cubes/inch), we get a prediction from the chart of some 116 degrees. To compensate for the valve angle, practice indicates anything from zero to 2 degrees, depending on the port downdraft. As for CR corrections, the same 0.75-spread per ratio over 10.5:1 seems to work fine. Fortunately, the LCA of a multi-cam, four-valve engine can be adjusted, and therefore computing the LCA beforehand is not so important. However, the LCA for a single-cam engine is fixed by the grinding process, and the LCA needs to be precisely determined up front. The chart’s prediction works well for a four-valve engine having an effective valve seat form achieved via a well-modified cylinder head. If the valve seat is of a typical factory form, the low-lift flow, although strong, is

typically and measurably less than a flow-developed form. This being the case, we find that standard heads, or heads employing standard valve and seat forms, tend to want a slightly tighter LCA than predicted here. Usually, 1- or 2-degrees reduction is sufficient to compensate.

Duration We have learned how to make a good determination of the overlap and the cam’s LCA for a given application. When the numbers for the overlap and LCA have been determined, only one intake duration figure fits those numbers. For instance, if, for a single-pattern cam, let us assume the overlap called for is 60 degrees and the LCA 108, the duration can only be 276 degrees. To arrive at the duration, when the overlap and LCA are known, we take the overlap (60) and divide by 2, add it to the LCA (108 + 30), and then double it. That’s 138 × 2, which equals 276 degrees. One aspect that all too often afflicts those with the racer mentality is what I like to call the “Stroker McGurk” syndrome. In so many cases, a racer works on the principle, “If some is good, then more must be better, and too much must be just right.” That comes into play far too often when cam selection is the job at hand. If you were buying a regular catalog cam, the chances are it would be a wider LCA than any cam computed from data in this book. Be aware that some cam grinders do this to protect the end user from their own desire to use too much duration. Rather than argue, they spread the LCA to keep the cam more civilized, but the customer incures a penalty and that is less output than would otherwise be the case. Exhaust Duration From what we have discussed so far, you can see how the duration is dependent on the overlap and LCA. Of course, we have not looked at the whole picture yet—among other things, the intake-to-exhaust duration. What is done here is to compare the intake and exhaust flow and determine how much exhaust timing is needed to adequately rid the cylinder of its contents. This is influenced primarily by the intake-to-exhaust flow ratio of the valves, but it is far from the fixed value often supposed by cylinder head porters. The CR and RPM strongly influence the optimum exhaust-to-intake ratio. Just for the record, when developing a head, the higher the intended

CR the less exhaust flow is needed in relation to the intake. A 10:1 CR may produce best results with the commonly used 75- to 80-percent exhaust-tointake flow. A 17:1 CR, however, can produce better results with a slightly bigger intake and a slightly smaller exhaust, about 65- to 70-percent exhaust-to-intake ratio. This can be made to produce better results because a high-compression cylinder puts out most of its power earlier in the power stroke. This allows the exhaust to be opened earlier and a smaller exhaust valve can, as a consequence, be used. This, in turn, allows for a bigger intake valve. When a cam is selected for a motor, it has to match the cylinder head flow characteristics, be they optimum or not. In other words, it makes no sense to install a cam ground to suit heads you don’t have. If the exhaust flow is too low, the cam needs to compensate in part by having a slightly longer duration and, as you would expect, vice versa. As the RPM band goes up the RPM range, so does the exhaust duration required. If a very high compression and a lot of RPM are involved, then the required exhaust duration can be increased (at the expense of exhaust diameter) by some 10 to 15 degrees more than the intake. This allows a larger intake to be used. Now that we have it nailed down, it’s worth looking at how duration affects power. Chart 10-6 shows torque and horsepower results from tests done on a 306 small-block Ford. This was equipped with out-of-the-box Dart aluminum heads delivering a CR of 10:1. The intake was a singleplane Edelbrock Victor Jr., utilizing a mildly modified 650 Holley carb. The solid flat-tappet cams tested all had the same lift (0.510). The shortest was 272 degrees at 0.020 tappet lift, the next longer was 285, and the longest was 300 degrees. As can be seen, a longer cam does not influence peak torque by any significant amount. Essentially, the same torque is produced farther up the RPM range and by this means produces more power.

Valve Lift: How Much? Under-valved engines like a lot of valve lift, and two-valve engines are generally just that—under-valved. Filling the cylinder is always the major problem, while dispensing exhaust is not so major. Given that, almost any two-valve pushrod engine needs as much valve lift as possible up to about 0.35 inch of the valve’s diameter. If we apply this to a typical pushrod

engine such as, say, a small-block Chevy, the lift called for with a 2.02-inch intake valve is 0.707 inch.

10-6. From these tests, it can be seen that torque does not increase with increased duration. What does happen is that low speed is traded off for high-speed output. These tests represent the worst-case scenario as the CR, which needs to increase as duration is increased, was held constant. Achieving that kind of lift is possible, but only under certain conditions. Assume, for example, we have a cam that is relatively short; something like a 260-degree hydraulic. Here, the short duration involved just does not allow enough time to reach the sort of lift that maximizes output. The tappet diameter limits the velocity that can be put into the follower on a flat-tappet cam.

However, the initial acceleration of a flat-tappet cam can be significantly higher than a roller cam. In fact, a roller cam’s strength is its ability to run to higher velocities than a flat-tappet. What this means is that although the roller cam is slower initially from the seat, it usually allows more lift than a flat-tappet cam. For a typical hydraulic valvetrain pushrod V-8, a roller versus a flat-tappet shows that, for cams below about 275 degrees, a flattappet delivers as much or more area under the lift curve than a roller. For hot street cams, the breakeven point is usually about 278 degrees. What this means is that, for short cams, anything that maximizes lift is going to be good for power. From this, it follows that high-lift rockers and setting up pushrod lengths to optimize geometry can pay off. From Chart 10-7, we can see that going to higher-lift rockers only minimally impacted low-speed output, but, from about 4,500 rpm on up, showed a very worthwhile increase in output. Our test example is a 350 small-block Chevy, on which lift is increased by swapping out the 1.5:1 baseline rockers for 1.6s, which produced a 19-hp improvement. And further increasing it with 1.7s added another 11 hp.

10-7. Here are test results on a 355-ci Chevy with a 10.7:1 CR, using some high-flow Dart Platinum heads in ported form. A 276-degree Comp Cams

flat-tappet hydraulic cam was installed but was spec’d on a 112-degree LCA, so that as lift was increased, the overlap would not become too excessive and skew results. Lift values achieved with 1.5:1-, 1.6:1-, and 1.7:1-ratio rockers were 0.474, 0.506, and 0.537 respectively.

If your goal is to optimize the valvetrain—it really is worth power—then time on a Spintron such as this one at Comp Cams is a good idea. It’s not cheap, but the results warrant its use, as many Cup Car teams have shown. There are, of course, diminishing returns from the use of ever-higher valve lift values, but the returns are good here because the lift used is still way below 0.700, plus that would deliver maximum power for the duration involved. Just for the record: The peak torque output rose from 430 to 447 ft-lbs on an engine that was re-equipped with a cam of the same profile on the intake and exhaust, but with a 108 LCA, 1.65:1 intake rockers, and 1.6:1 exhaust.

Another aspect we have to consider, when attempting to maximize lift, is that it involves increased valve accelerations and, thus, higher loads. More lift equals less reliability. Therefore, the engineering has to be much better. These days, building an all-out-maximum-performance valve-train, especially a pushrod one, almost certainly involves using a Spintron somewhere during the process of development. While we are on the subject of lift, it is worth pointing out this: As rocker ratios and, consequently, valve accelerations increase, the LCA needs to be spread slightly from that shown on the chart. If the LCA was optimal on, say, 1.6 rockers, then a switch to 1.8s can mean a spreading of 1 degree to restore the situation to optimal. One last point is that exhaust-valve accelerations are far less critical than intake. As a result, a lower-ratio rocker often suffices, especially on a dualpattern cam (more exhaust duration than intake).

CHAPTER 11

VALVETRAIN: THE PHYSICAL BUILD By now, I am sure you have a good idea of both the complexity and importance of the valvetrain, as this is the third consecutive chapter dealing with it. What I intend to focus on here is the practical application of what I have so far explained. At this point, you should know what events and lift your cam and valvetrain should produce to meet the application requirements of your engine build. Although that is great progress, the job is far from done.

Reading Cam Spec Data Cam profiles are rated at a number of degrees at a certain lift. Depending on the type of follower to be used, cam profiles are rated at differing lift values. For instance, a hydraulic cam is rated in terms of degrees of tappet lift at 0.006 and 0.050. Some cam companies, such as Lunati and Comp Cams, also quote the cam’s degrees at 0.200 lift. Using these numbers, we can subtract the degrees at 0.050 from the degrees at 0.006 to determine how “fast” the profile is. The smaller the number between these two lift points, the greater the tappet acceleration and the higher its speed at the 0.050 point. This is known as the cam’s hydraulic intensity. The 0.006 measuring point is about the amount a good hydraulic valvetrain collapses due to component flexure, lifter leakage, etc.; this means the degrees quoted are close to the duration seen at the valves.

Knowing what is required in theory and applying it in practice are two different things. In this chapter I will look at the practicalities of assembling a high-performance valvetrain. In much the same way, figures for a solid-tappet cam are quoted, but the primary lift mark is at 0.020 instead of 0.006. If all lash values, initial profile acceleration rates, and rocker ratios were the same, we could use these figures to determine the duration that exists after lash has been figured in. For some reason absolutely mystifying to me, the cam companies do not give us the duration at lash on a solid lifter. On top of that, many street profiles have the tappet lift quoted at 0.015 inches. What this means is that you need to make a guess as to what the real duration of a typical solidlifter cam is at lash. If I am perusing a catalog, I make a provisional assumption that most cams rated at 0.020 are between 12 and 15 degrees longer at lash. Bear this in mind when comparing a solid profile with a hydraulic.

When the cam spec you require has been determined, you will find it much easier to make a choice from the cams listed in a manufacturer’s catalog because they could well have an off-the-shelf grind really close to what your engine needs.

Assembling a really functional valvetrain is more than just selecting parts almost at random. Dynamic compatibility is paramount. The first move toward this goal is to select a top-quality valve spring.

Valvesprings It may seem less than obvious at first, but a valvetrain can be no better than the valvespring that controls it. This makes the selection of the valvesprings for your engine of great importance. Just having the cam manufacturer’s recommended seat preload and over-the-nose spring poundage is only part of the equation. Since we are on the subject, the term so often used to describe the preload on a closed valve is “seat pressure,” but that is not the correct term. Pressure is pounds per square inch (like you put in tires) or load per unit area. We are dealing with load here, plain and simple, so the correct term is: “seat preload poundage,” or simply seat poundage.

A big-block Chevy has a big, heavy valvetrain, so things need to be done right. The first two builds here used AFR as-cast heads at 10.5:1 CR. The blue curve is for a 280 Lunati solid street roller. The black curve is for a Comp Cam circle-track grind with the same 280 degrees duration at 0.020 as the Lunati street roller, but with 2 degrees more at 0.050, 4 degrees more at 200, and about 0.050 more lift at the valve. The red curve is for a set of CNC-ported AFR heads, a 12:1 CR, and a Sniper intake. The cam was 298/309 at 0.020 and 0.800 net lift after lash.

A heavy blow to the top of the spring momentarily compresses the top coils. This coil compression then travels back and forth along the spring’s length until it dies out. Note how the poundage exerted by the spring momentarily drops when the coil reflection occurs. A spring’s stiffness is expressed in pounds per inch. A spring that has, say, a 250-pounds-per-inch rate delivers a force of 250 pounds for every inch it is compressed. Initially, we are going to be looking for a spring package that delivers the seat and nose poundage called for by the cam manufacturer. If we want to build as good a valvetrain as possible without resorting to the expense of hiring a Spintron, then the next step is to find a spring package that has the least mass. The way to look at this is that the spring has to use a certain amount of its valvetrain-closing force to return its own mass to the valve’s closed position. If the spring is heavy, then it has less remaining force to operate the rest of the valvetrain. There is also a little-known but immensely important fact known as the “resonant frequency.” When the principal valvetrain harmonic hits this frequency, it drives the spring into what is called “spring surge.” This uncontrolled motion of the spring is usually a giant power killer. Understand that valvetrains have components with their own resonant frequencies and, as such, there is almost always a cam profile motion frequency that is at or a multiple of one or other component.

There are two ways in which we guard against these spurious vibrations from becoming a problem to motion control. The first is to damp the spring, and the second is to use a spring package that has as high a natural resonant frequency as possible. That means selecting a spring pack with the minimum mass. The problem is, most cam companies, with the exception of Crower, don’t put the weight of the spring in their tech sheets. Your other alternative is to go directly to the spring manufacturer. PAC springs even go as far as quoting the resonant frequency of their springs. This company is a source I can recommend.

A pushrod vibrates laterally. For the most part, this needs to be at as high a frequency as possible (via larger diameter and thinner wall thickness). Although, with careful tuning, this lateral flex can aid power by means of valve toss.

Not only is a spring required to contain and damp-out the frequencies at which it may vibrate, but to dampen the vibrations inherent from other components in the valvetrain. Of these, the pushrod can either be a source of severe hindrance or a power advantage. Where rules limit the tappet diameter (such as NASCAR’s premier Sprint Cup series), the amount of lift that can be put into the valve can be somewhat restricted. By tuning the pushrod flex, the valvetrain can, at high RPM, be made to “toss” the valve in a reasonably controlled manner. This extra lift is said to be worth as much as 20 hp. There are valid occasions when valve toss can be used in other classes. One example that comes to mind is when valve lift is limited by the class regulations. If the intake valve can be coaxed to 0.050 higher lift, there is most certainly a power advantage, as long as it does not subsequently bounce off the seat. Such valvetrains are somewhat specialized in nature and should be regarded as difficult to build without the precise input of Spintron data. Also, it is not a good idea for a roller valvetrain to go into any motion that deviates from the cam’s profile. Although we tend to think of a valvetrain motion as being in a controlled manner all the way up the RPM range, until it reaches valvetrain crash speed, this is not always so by any margin. Spring surge mostly occurs at a speed substantially below ultimate crash speed; and when that happens, it causes a dip in the engine’s output at that RPM. The cure is to change the spring preload, natural resonant frequency, or damping. If the surge is minimal, shimming the spring to a higher seat preload can be a fix. The next option is a change to a lower-mass higher-frequency spring, and the last option is to change to a spring pack that has higher damping rates. We have not talked about damping yet, but the normal method of dampingout spurious vibrations is to introduce inter-spring friction either by means of a flat-wound damper between the inner and outer spring or reducing the friction of each spring (of a pair) with itself.

These two springs exert virtually the same seat and nose poundage, but the one on the right from PAC is made of a superior material and weighs only 65 percent of the one on the left. This increased the RPM to valvetrain crash from 6,100 to 7,150 rpm. Since the turn of the millennia, beehive springs have become very popular with the performance community, and not without good cause. This type of design has the advantage that the changing pitch and diameter of the coils causes it to have significantly less of a clear-cut resonant frequency. As the spring is compressed, it seats one coil down upon the one below it. This causes the rate (stiffness) of the remaining active coils to increase, and the mass involved to decrease. By the time the spring is set to resonate at one frequency, the situation changes and another frequency comes into play. Since the spring goes through these frequency changes so fast, it has little effect on the spring’s overall performance.

Before assembly, each spring should be measured and then shimmed at installation to give the desired seat preload.

Here are the two principal types of motion that occur when the valvetrain no longer follows the dictates of the cam’s profile. Valve loft, or toss, can help power, but valve bounce is a power killer. This sounds good in theory, but the real question for an engine builder lacking a Spintron machine is, “Does this really work?” The answer is yes and, these days, I use mostly beehives on any valvetrain for which I do not have time to get Spintron data. By going with the beehive spring, I am putting myself as far theoretically from valvetrain issues as possible. My advice here is: Regard a beehive spring as your number-one choice in the absence of really top-notch expert advice to the contrary. Some of the data from Spintron tests I ran with the engineers at Comp Cams demonstrate just how well this type of spring can work. This type of spring is especially effective with a hydraulic cam; a lower preload can mean less hydraulic collapse during operation and less of a chance of lifter pump-up.

Here is what valve toss and valve bounce look like on the scope while being run on a Spintron test rig. Valve bounce within the RPM range used has to be avoided if output at that point is to be preserved.

Rockers produced from aluminum or stainless steel have a degree of internal damping. Such rockers, when paired with a functional spring, such as the beehives shown here, can produce a highly functional and well-behaved valvetrain.

Because they rely heavily on the changing of the natural resonant frequency to retain control, beehive springs are at their best when run close to coil bind at full lift. I have had excellent results in valvetrain dynamics by running these springs to within 0.015 inch of coil bind.

A beehive spring’s function is much more complex than its regular counterpart. Top-grade beehives, such as this PAC item, not only have the obvious variable-wound coil but are made of an ovate-section wire to more evenly distribute stresses within the wire.

These two springs were run on Comp Cams’ Spintron. The regular spring delivered higher seat- and over-the-nose poundage by about 12 percent, yet the beehive ran to more than 1,000 rpm higher before losing control. Because pushrods are intrinsically tied in with the spring’s performance, I’d like to mention a few things about them here. These days, intense Spintron testing has made pushrod design a highly focused science. Pushrods are not simply a steel tube, made of some reasonable grade of steel, with a couple of ends attached. The material, heat treatment, wall thickness, OD, and form (tapered or parallel) can make a significant difference between a so-so pushrod and a really good one. Unless you are doing little more than a rebuild with a mild cam change, my advice here is buy a good brand because it may well save you some difficult-to-diagnose dynamic valvetrain problems.

Tappets The simplest tappet or lifter is the solid variety for a bucket-type overhead cam valvetrain. So long as the tappet is light, resists wear, and is of adequate diameter, there is little else to concern us. With the removal of the high-pressure lubricant ZDDP from ASI approved oils, we have to assume the need to minimize the possibility of wear much more seriously than in the past. (Note: You don’t have to use these non-ZDDP oils, so here I recommend using specialist oils such as discussed near the end of Chapter 12.)

It helps to understand the mode of function of a flat-tappet lifter. First, they are not flat. A flat-tappet typically has about an 80- to 100-inch radius on the face. Second, there is usually a taper on the cam lobe. Finally, with the cam in the installed position, the cam lobe does not wipe centrally across the tappet, but to one side. When working as intended, all these factors cause the lifter to rotate over half a turn during the lift event. This means the rubbing speed of the cam profile on the lifter is only a fraction of what it would be if no rotation took place. So that you understand the importance of this rotation, be aware that a regular cam won’t last 100 miles without rotation taking place. As added insurance against wear, some cam companies such as Comp Cams and Lunati offer a nitride hardening service for flat-tappet cams. My advice is: Pay the $100 extra and have it done— you won’t regret it.

Even with 12 pounds or so less seat preload, the beehive spring gave better control, plus an additional 1,000 rpm to crash speed.

Although less applied spring force from the beehive cuts rocker flex, most of the spurious valve-lift motion seen here is due to pushrod flexure releasing its energy back into the valvetrain.

Here is the valvetrain I favor for hydraulic applications. Note the small steel beehive retainer is as light as a titanium regular spring.

The taper across the lobe width provides the means for vitally important rotation onto a flat-face tappet. Check this out: It should be about 0.001. If the cam has no taper, call your supplier and ask why not. The chances are you need the cam replaced. When building a flat-tappet engine, be sure the tappets are free to rotate in their bores and that you use the high-pressure break-in lube supplied with your cam by the cam company. Fail here, and your cam and tappets fail right along with you. Flat Tappet or Roller? As mentioned earlier, a flat-tappet cam can have higher acceleration rates on initial opening and closing. Where it can fail is in terms of the maximum velocity that can be imparted to it. The maximum velocity is governed by the tappet diameter. The bigger the diameter, the greater the maximum velocity can be. A roller cam’s acceleration is limited by the side thrust seen as the tappet is lifted. The size of the base circle and the pressure angle formed on opening is the controlling factor. The bigger the cam’s base circle, and the bigger the roller, the lower the pressure angle becomes, thus allowing an increase in initial acceleration. Where a roller tappet scores is that it is not restricted in the same manner as far as velocity goes, and much higher values can be achieved so long as the duration is sufficient to allow it.

Money spent on quality tappets (lifters) is money well spent. These armored items are, when used as intended, virtually bullet proof and as such can go a half-million-plus miles.

Supplying oil to the interface between cam and lifter is a good choice for increasing the reliability and life of the tappet and cam. Comp Cams offer a lifter-bore grooving tool that spills oil from the lifter oil galley onto the face of the cam. The whole job is easily completed in 30 seconds, and it works well. Also you can get solid tappets with a pin hole in the face to supply oil to the cam/tappet interface. Solid rollers are intrinsically trouble free, so long as they are not overloaded, but the same cannot be said for hydraulic rollers. Distortion of the body from side loads appears to open up the internal clearances to the extent that they collapse all too easily. The extent that this can happen is far greater than is often supposed. While doing some Spintron work testing springs and hydraulic rollers, I found that OE rollers could collapse and reduce valve lift as much as 0.100!

This cam was installed with no heed to the guidelines offered in this chapter. It has less than an hour’s running time on it!

The roller diameter and the cam’s base circle diameter limit the amount of the acceleration on a roller tappet. A pressure angle (arrow A) of about 28 degrees is normally considered about maximum but, if a short-duty life is all that is needed, this can go up to about 33 degrees.

Timing Gears

Timing-in the cam for a standard rebuild usually only involves aligning the timing dots on the gears. All too often, performance builds are done like this in the belief that this is close enough. If you thought that this might be the case, see photo 11-1, which shows just how far off the stock gears can be. Although I saved this set to show an extreme case, it is certainly not unusual to see the timing 2 to 3 degrees off what may be needed for best power.

This style of roller tappet, with the horizontal spring-loaded link bar to prevent rotation (only available for small-block Chevy and from Crane Lunati), has a significantly lower reciprocating mass than the conventional vertical link bar design.

This Endurex solid-roller tappet from Comp Cams, for a big-block Chevy, represents a typical roller-design style with the vertical link bar.

The billet body hydraulic Lunati rollers seen here are built to far closer tolerances than the OE rollers. All too often, the use of OE-style hydraulic rollers with an aggressive cam profile can cost in excess of 50 hp; although more expensive, the fix is to use these Lunati rollers. The increased power potential more than justifies their extra cost. Any serious engine build should involve timing-in the cam accurately. This should happen twice; the first time during a dummy build to establish that, when the cam is where it is supposed to be, there is adequate valve-topiston clearance. The second time should occur during final assembly. Most cam companies call for a minimum of 0.100, but that, to a large extent, is a failsafe deal. With the benefit of spin testing and the use of premium parts and profiles I have run valve-to-piston clearances as little as 0.016 on smallblock Chevy engines turning upward of 8,000 rpm. Even the experts at the cam companies whose parts I used were somewhat in awe of the success with such close clearances. Back in my British Touring Car Championship era, I had the fastest car in the class, and the 10,500 rpm that my pushrod engines ran to still only had 0.030 intake-to-piston clearance. I am quoting those numbers just to make the point that spin testing can be a great advantage in optimizing the dynamics of the valvetrain.

Cam Timing A great deal of emphasis is put on the timing of the cam into the engine. The more accurately the cam is spec’d-out, the greater the need is to do this. The intake lobe centerline method is the most common means of quoting the cam’s called-for timing into the engine. Without going into great detail as to why, this is by far the best technique to adopt. I also don’t intend to go into detail as to how this is achieved. If you buy any cam-degreeing kit from any cam manufacturer or go to their web site, you find a fully detailed description of how it is done.

11-1. From these timing marks, you can see that the cam would be timed a whole tooth out. On a nominally 300-hp V-8, that would cost as much as 50 hp.

If a roller cam is being used in a block that was otherwise equipped with a flat-tappet design, there is often a need to control cam end float with a solid or roller thrust button (shown here).

The question I am posing is: Are you aware of what the correct cam timing should be and why? Sure, you can just take the cam grinder’s word for it and maybe that should be your primary choice in the absence of any other concrete data. However, it won’t hurt to just discuss what factors may make the optimum cam timing other than that recommended by the cam manufacturer.

Although it may at first seem unrelated, the ability of the crank damper to damp-out torsional crank vibrations is key for a smooth-running, problem-free valvetrain. Without a damper, crank vibrations are superimposed directly onto the cam’s motion.

Here, race-winning engine-builder Dusty Kennett times-in the cam on this 8,000-rpm, small-block Chevy. Getting optimal results depends on knowing what they should be and what factors may affect where the cam is ultimately timed-in at. Before we go any further, I should explain how advanced and retarded cam events are defined. If a cam is on, say, a 110-degree LCA, then, if it is timed-in with number-1 intake reaching full lift at 110 degrees after TDC on the inlet stroke, the cam is said to be timed “straight up.” That is, no advance or retard. If all the cam events happen earlier, the cam is advanced. The amount advance is defined as the LCA minus the intake centerline (that’s the point at which full lift occurs—ATDC.) Say we have a cam on a 108 LCA timedin to 4 degrees of advance. This means the intake centerline is at 104 degrees. If it is retarded 4 degrees, it is at 112 degrees.

This split timing cover for a small-block Chevy allows access to the cam to make timing changes without having to remove the crank damper and drop the sump.

To make sure the cam timing is optimal, I make a great deal of use of adjustable cam timing sets. One of my favorites is this unit from Jesel. This particular example is for a small-block Ford, but the company also makes them for several other popular Ford and Chevy engines. For many applications, such as two-valve pushrod V-8s, the cam manufacturer calls for 4 degrees of advance, and in most cases this works. But at the same time, it does make assumptions. For instance, let’s say the machine shop guy was brilliant at cutting intake seats that had considerably more low-lift flow than the normal three-angle valve job delivers. In this instance, the flow in relation to the piston and the exhaust valve appears to the cylinder as if the cam were too far advanced. For this to work best, the cam may need to be pulled back 2 degrees or more.

Adjustable cam drives, such as this unit from Cloyes, represent a costeffective means of adjusting timing in a relatively speedy manner, especially when used in conjunction with a split timing cover.

Milodon is noted for its top-of-the-line gear drives. These feature all the timing and control assets of a belt drive but use gears instead of belts. Such systems are popular with alcohol burners and GMC-style blower applications. Some of my own seat work produces low-lift flow equal to a valve of significantly bigger proportions, and as a result the cam has to be timed-in as much as 2 degrees retarded. The difference between the usually recommended 4 degrees advance and the optimal 2 degrees retarded, with my heads, is the difference between 575 hp and more than 600 hp. So accurately timing the cam, is worth the effort, but the only way to be absolutely sure it is optimal is to do the final setting on the dyno. One last point on the subject of cam timing: Why do better spec’d cams need to be timed more accurately? The answer is less than obvious. The principal reason is that if there is an error in the exhaust events one way,

and an error in the intake the other way, we find that, as the cam is, say, advanced, the intake improves while the exhaust gets worse. The result is that the timing is not that fussy if one event is improved at the expense of the other. If the valve events were perfect, then any change in advance causes both the intake and exhaust event to get worse and makes power drop-off much faster. Now you know!

This rocker setup may look heavy and bulky, but appearances can be deceiving. In practice, this type of rocker made a very dynamically stable valvetrain able to impart almost 0.700 inch of lift up to speeds well in excess of 8,000 rpm.

Rockers In this day and age, you might wonder why any engine has rockers, or even pushrods for that matter. But the reality is that the American V-8 is still a dominant factor in the performance world. Its sheer size works for us —that and the fact that, over the past 50 years, the performance industry in the United States as a whole has become really good at finding horsepower. Add to this its basic simplicity and, as a result, we still have pushrods and rockers. Like most things, there is more to the design of a rocker than meets the eye. As I alluded to previously, a rocker needs to not only amplify and transmit the required motion to the valve, but also damp-out high-frequency vibrations as far as possible. Although rockers made of steel can perform well, aluminum does have a moderate advantage here by being easy to machine.

Rocker ratio geometry may look as simple as dividing R2 into R1, but there is much more to it than that. As the pushrod pickup point moves around in the direction of the green arrow, the initial opening ratio gets higher for a faster off-the-seat opening. That’s usually good for extra power.

Pushrod length must be selected so that the sweep of the rocker across the tip of the valve is reasonably well centralized, as shown here. The first thing to consider when selecting rockers is the ratio needed for best results. Over the years, trial-and-error testing of many rocker designs has led engineers to the conclusion that ratios between about 1.4 and 1.7:1 are effective and do not present any great manufacturing or design challenges. But the dyno shows that valve lifts in the region of 1 inch or more can pay off, and this has led to the development of ratios as high as 2.1:1. For a pushrod two-valve engine, we can reasonably say that the greater the rocker ratio, the higher the power potential. But anything exceeding about 1.8:1 starts to get a little fussy in the way of the hardware to support such a ratio. Everything in the valvetrain such as pushrods, springs, valve mass, tappets, etc., starts to demand significantly higher sophistication.

This rocker setup utilized a 1.75:1 ratio on the intake and 1.65 on the exhaust. A split of 0.1 of a ratio between intake and exhaust usually works because of lower exhaust-valve accelerations but, with a little more duration, have been shown to work well for high-rpm applications.

By making the stem hollow, this Ferrea intake valve drops from 119 to 101 grams.

When testing to determine the results of an increase in rocker ratio, do not overlook that the overlap triangle is increased with the increase in ratio. To compensate, the LCA needs to be widened to truly establish what the higher ratio actually does for output.

Polylocs in the lash adjuster have a habit of coming loose. The use of a stud girdle on a stud-mounted rocker system not only prevents that but also provides added stabilization of the valvetrain.

Having all the right components in the valvetrain is a good start, but the job does need to be finished by correctly lashing it. Here, a little fine tuning of the lash can possibly add just a few extra horses. Usually having the intake slightly looser than the exhaust works best, but the dyno provides the final proof.

Here is the result of a Spintron test on three weights of intake valve in a pushrod valvetrain. On the left is a 119-gram valve, in the center is a 101-

gram hollow-stem valve, and on the right is a 89-gram titanium valve. Saving weight in the valvetrain is worthwhile, especially at the valve. A titanium retainer instead of a steel one is typically worth 100 rpm in a pushrod valvetrain. Hollow-stem valves like those from Ferrea are more than halfway to a titanium valve, but at less than half the cost.

CHAPTER 12

THE SHORT-BLOCK ENGINE We have now arrived at the long-block assembly and, for the production of power, our principal goal is to minimize friction. But that is something of an oversimplification of our mission statement. There are many other practical aspects we have to address, beginning with the engine block.

The Block The block is, in effect, the case that has to contain all the loads brought on by the development of the power we are seeking to build. If you are building an engine for any serious competition, the first move should be to sonic-check a selection of blocks. This allows you to throw out any that may have thin sections in high-stress areas and to sort out the best of the selection at hand. After you choose a suitable candidate, the main bearing housings need to be checked for alignment and size. Here, the housing diameters need to be sized to ensure the bearing crush is slightly on the tight side, so as to avoid any chance of a spun bearing.

The block and rotating assembly are fundamentally important components of the engine. In operation these components are subjected to the cylinder pressures that are turned into usable rotational power. Understanding the vital components and figures will mean not only good power but also reliable power production. The bores then need to be considered. The boring equipment in many machine shops locates off the old bore, and thus replicates any error there may have been with the bore’s alignment over the mains. The best way to handle things here is to bore on a machine that locates on the main bearings, and reference from there. Also the block faces need to be trued-up square and parallel with the mains. For certain classes of racing, other jobs may also be necessary. One that comes to mind is accurately locating the tappet bores and resizing them to

the largest allowed by the rules. Bore finish is also very important, and it’s mostly dictated by the type of ring material used. Unless you have positive experience to the contrary, follow the ring manufacturer’s recommendations to the letter.

The Rotating Assembly Depending on the extent of the quest for power and the budget available, there is much that can be done in the pistons, rods, and crank departments. Before we delve too much deeper into short-block tech, it’s worth taking a look at all the parameters that define the bottom end assembly, and factors that we have some control over.

Here are the bores on one of my 5.0 race engines. Size and finish here is everything. Get the best job you can afford because it pays dividends.

The block is the primary component for any engine. The block needs to be strong and machined to fine tolerances if winning results are to be achieved.

Sonic testing a block not only saves you the cost of a wrecked engine if a section was too thin, but also helps to find that 1-in-50 block that can be bored as much as an extra 0.060 inch oversize.

The greater the success at finding power from breathing modifications, the more critical the strength and geometry of the rotating assembly components becomes. Here is one of my Scat stroker rotating assemblies. This stretches a 350 Chevy to 383 and is good to 7,700 rpm on the street.

The geometry of a crank/rod/piston assembly is a very effective means of converting high-pressure gases to rotating mechanical power. If you

intend to spec-out and build your own engines, you need to know the relevance of the following dimensions (red arrows): Deck height (1) is the dimension from the crank centerline to the head face of the block. Crank throw (2) is the radius the rod journal sweeps out as it rotates around the main journal centerline. Crank stroke (3) is twice the crank throw and represents the amount the crank moves the piston up and down the bore. Rod center-to-center length (4) is most usually referred to as the rod length. Piston compression height (5) is sometimes also called the pin height; this dimension refers to the distance between the center of the wrist pin and the top surface of the piston. Swept volume or cubic inch displacement, or CID, (6) is the amount of air the cylinder is capable of drawing in as the piston moves from the top of the stroke to the bottom. You also need to know about these components (blue arrows): block deck (1), cylinder bore (2), piston (3), wrist pin (4), ring belt (5), connecting rod (6), rod journal (7), crank counterweights (8), and main journal (9). One of the primary considerations when designing an engine from scratch is the bore/stroke ratio. If maximum power/torque is the primary criteria then, for a given displacement, a big bore and a short stroke gets the job done—it allows bigger valves to be accommodated. However, for a production engine, emissions and packaging (engine length) usually dictate longer strokes and smaller bores. When Formula 3 was a pushrod one-liter class breathing through a single 28-mm restrictor bore/stroke ratios, winning engines were typically in the region of 1.8:1 (bore 1.8 times the stroke). The short stroke allowed for very high RPM. The last one I built (went into my 1-liter Ford Anglia) turned a useful 11,000 rpm. A modern production engine such as, say, the Ford four-valve 5.4-liter Modular Motor, that powers the 500-hp Shelby 500 Mustang, is at a 0.85:1 ratio. Cranks are essentially available in three grades: cast, forged, and billet. Cast is the least expensive, billet is the most expensive, and forged is in between. Many production vehicles employ cast-iron cranks because of ease of manufacture and wear resistance. These won’t take either the outright power or abuse forged ones will, but they are usually okay for a reasonably strong street performer. A forged crank is a step up the ladder and, for more money, provides greater strength. Although factory cranks are usually of lesser-grade steels, the steels most commonly used in aftermarket cranks are 4130 and 4340.

Depending on the cost of the crank, these can be had with surface hardening, which gives a very good wear life. Last on the list are billet cranks. These are made from a block of steel because there is no appropriate forging available from which to machine the desired design of crank. And they can be very expensive. With aftermarket cranks becoming a lot less costly, it becomes very practical to increase the stroke of an engine just to get more displacement. Buying a new stroker crank also means you have a stronger crank, with 100 percent of its fatigue life still available. If a stroker build is in the cards, the next step is to choose your brand of crank wisely. I have dealt with Crower, Scat, Lunati, Callies, K1, and Winberg to good effect and can recommend these companies. The downside of an increased stroke is that the increased piston speed means reduced RPM potential. If we are building just to make the most power from the engine rather than to any race regulations limiting displacement, then the advantage of more cubes from a longer stroke more than offsets the disadvantage of reduced RPM.

My dyno testing has indicated that, for a typical wet-sump V-8 having a 3.75-inch stroke, windage losses can be measurably cut. The flat-face counterweight crank (top) produced 7 hp less at 7,000 rpm than the Scat aero crank (bottom).

Factory V-6 or V-8 cranks often are externally balanced by means of a counterbalance mass in the damper and flywheel. To avoid this lessdesirable form of balance, the big-end journal can be hollowed out and heavy metal added to the counterweights.

After 100-percent surface machining, this crank was shot peened. This puts a compressive stress into the surface of the crank and has the effect of increasing its fatigue resistance.

This is an internally balanced (three heavy-metal slugs were used) 4.1inch-stroke crank for a 635-hp, 425-ci, drag-race small-block Ford. This relatively simple and inexpensive build used a limiting mean piston speed of 5,200 ft/min, which equated to a red line of 7,600 rpm.

If the crank is to have any chance at long-term survival, a damper that functions in the RPM band used is vital. The BHJ unit seen here is very popular with top professional teams.

Put on the crank degree pointer for ignition timing before the heads are installed, so that the TDC position can be accurately set. Just how many RPM the rotating assembly can handle depends on the crank stroke and the strength of the connecting rods that will be used. A good cast-steel crank and a decent set of rods should be good for a mean piston speed of 4,500 feet per minute. To figure that in terms of RPM, take the stroke in inches, divide it by 6, and then divide that answer into the limiting mean piston speed. In the example we would, for a 3.75-inchstroke crank, have 3.75/6 = 0.625. If we now divide 0.625 into 4,500, we get 7,200. That is our limiting RPM. If the stroke had been of 3.5 inches, then the RPM for the same 4,500-ft/min average piston speed would have been a shade over 7,700 rpm. If top-quality race parts are used, the limiting mean piston speed can be increased to 5,000 ft/min and, for a drag-race-only application, even higher. As of 2010, a front-running 500-inch Pro Stocker can turn more than 10,000 rpm. With the typical stroke of around 3.7 inches, this equates to a mean piston speed of a blistering 6,200 ft/min. This piston speed may be fine if you are not expecting to put more than a couple of hours or so of race time on the parts involved.

Connecting Rods It’s possible to fill an entire chapter on the subject of connecting rods, but I do not have the luxury of space to do that. Here, I need to cover the

important basics, starting with the way we discuss rod proportions. Because engines vary so much in size, simply quoting the rod’s center-to-center length tells us little about its geometric proportions. The most important element of a rod is its length in relation to the crank’s stroke. This is called the rod/stroke ratio. What is used can have a significant effect on the engine’s RPM capability and its output. Illustration 12-1 shows how the rod/stroke ratio is defined. To come up with this the stroke, (A) is divided into the rod’s center-to-center length (B). If, in this case, the stroke was 3 inches and the rod 6 inches, the rod/stroke ratio would be 2:1.

12-1. An engine’s rod/stroke ratio is defined as the rod length (B) divided by the stroke (A). So what is the optimum rod/stroke ratio? If you ask around, you hear all kinds of theories as to what it should be and what is optimal. Most of what you are likely to hear is disinformation. The most likely piece of info you might get is that a long rod/stroke ratio is good for high-RPM power, and a short one is good for low-speed torque. The first part may be so, but the second, concerning short rods, is highly suspect. What defeats the situation is a combination of rod angularity and friction.

In simple terms, the longer the rod is, the more mechanically efficient the linkage system becomes. By the time ratios get to the 2:1 mark, there is little to be gained from further rod-length increases. There is one saving grace for the shorter rod ratio: The higher the compression ratio goes, the less negative effect there is with a short rod.

The shorter the rod becomes in relation to the stroke, the greater the rod angularity. When the gas pressure is applied to the piston, a higher rod angularity causes the piston to be pushed into the bore with greater force; this increases the losses to friction.

A high-strength, lightweight connecting rod is a prime requirement for any high-output build.

The Mahle piston shown here carries many high-tech features that seek to cut weight and friction while increasing strength and performance. This is paired with a sportsman-style rod. Made from the same material as more expensive rods, it has forged, non-machined surfaces to save cost.

Pistons Pistons can be had in three distinct material groups. These are, in order of preference for a high-performance engine: high-silicon aluminum casting, hypereutectic casting, and forging. Pistons for a regular street application need to be inexpensive to manufacture, exhibit long wear life, and run quiet. Casting a piston in the first place allows for the production of a cheap blank. Alloying the aluminum with a substantial amount of silicone makes the casting harder while reducing the expansion coefficient. Regular cast pistons take just so much abuse and then fail, usually catastrophically at that. The next step up the ladder is the hypereutectic cast piston. The term hypereutectic means that there is more of an alloying element present than will disolve in the parent metal. In this case the principal alloying element is again silicone, but in such proportion (11+ percent) that there are free crystals of silicone present. This makes for a harder piston and, given suitable section thickness, a reasonable strength. Depending on the bore size (smaller bores dissipate piston-killing heat faster) a hypereutectic piston can be used up to normally aspirated power levels of 100 hp per liter (1.65 hp per cubic inch). Such power levels need to be reserved for engines with bore sizes up to about 70 mm. For the 4inch-plus bore size of a V-8, power levels need to be contained to 70 or maybe 80 hp per liter to retain the structural integrity of the piston.

Street-oriented pistons often have a slot instead of the preferred drilled holes as seen here. The slot acts as a heat dam to cut skirt expansion, but weakens the piston. Last on the list is a forged piston. Most manufacturers work with two different alloys: one with a lesser expansion rate and which is harder after heat treating, and a second alloy for race purposes, which exhibits slightly greater expansion and less hardness, but is tougher. As may be expected, the physical form of the piston for a high-output application is important. When selecting a piston, you need to look for a well-supported skirt and pin boss along with drilled oil-return holes in the oil ring groove, rather than slots. The compression rings need to be placed as high as possible, consistent with the application. Go too high, and the rings and top-ring land suffer thermal degradation (burning and melting). Nitrous or forced induction calls for a lower top-ring placement than normally aspirated applications.

This Wiseco race piston carries many performance-enhancing features. These include high ring placement, thin ring sections, a shorter/stiffer/lighter wrist pin, thermal-barrier coating on the crown, and anti-friction skirt coating.

Look carefully at the skirt on this DSS race piston; the X-groove across the skirt face is evident. This grooving improves oil-control properties and allows a small amount of metered oil to the skirt face for friction control.

In an effort to increase rigidity and strength while minimizing weight, a modern race piston employs a strutted and braced design such as seen here. Also note how close the pin bosses are to the center of the piston, allowing for a shorter pin to be used.

12-2. The Total Seal top compression ring is a two-component design. The gaps of the major and minor ring are positioned on opposite sides of the piston. When gas pressure gets behind the major ring, it presses the major and minor rings onto the cylinder wall. Since the rings seal against each other and the gaps are opposite, this leaves no direct path for leakage to occur.

Rings There are many piston ring manufacturers. Each has a wide and often bewildering range of rings from which to choose. As you might expect, cheap rings are giving up something compared to what you may get if you pay more. Usually low cost means a shorter life, but that should not imply that any ring set that is not expensive is substandard. Many good ring sets cost less than you might expect because they are produced in such huge quantities. To ensure a reasonable life expectancy, make sure the rings are a suitable, heat-treated steel alloy or have a wear-inhibiting, moly-plasma spray coating. If the application is nitrous or supercharged, a ductile iron or, better yet, steel ring is needed to combat temperatures and shock loading. If you are going for high-end race rings, the titanium nitride steel rings (gold in color) are about as wear resistant as you can get. With ring styles, the less mass the ring has, the better. For a typical 3¾inch (or bigger) bore, ring widths of 1/16, 1/16, and 3/16 have been common for replacement pistons. But the need to meet mandated fuel-

economy standards has brought about the development of ring packages that seal better and also have lower friction values. Both Mahle and DSS have very affordable ring packages that feature lighter rings with thinner sections. The point here is: The thinner the ring, the more effectively it works in a high-performance engine. With today’s titanium nitride coatings, ring wear is no longer a challenge so, other than cost, rings of a very narrow cross section are not a problem. The 1-mm, 1-mm, 3-mm ring pack is becoming the choice of many engine builders, but section thickness is far from the only factor to take into account. There are various styles of top ring, of which the most common is the flat-face type. Probably the most complex is Total Seal’s top rings, which function as shown in Illustration 12-2. I have done much back-to-back testing of this type of ring, and it does do what is claimed of it: a virtually total seal of the cylinder. The advantage of this type of ring is that it nearly provides a 100-percent seal with a working-end gap of as much as 0.060. Such a wide gap puts it well out of the range of possible seizures due to thermal expansion causing the gap to entirely close, making the ring nearly a press-fit in the bores. Because it represents a failsafe deal and because results are consistently topnotch, I use Total Seal top rings in all but the cheapest builds that I do. Although it is termed a compression ring, the function of the second ring is not to seal combustion pressures so much as to act as the first line of oil control. The ring of choice here should, in most cases, be of a Napier Scraper design. Here the tapered face of the ring tends to ride over the oil on the way up the bore and scrape it off on the way down. The most common type of oil-control ring is the multirail-and-expander design. Introduced in the early 1960s, its effectiveness has been proven in many millions of applications. Part of the ring’s oil-control function is due to the outward radial load it exerts on the bore. This, of course, is a frictional loss, so it must be minimized. At this stage, crankcase vacuum comes into play. By pulling a partial vacuum on the crankcase, the need for so much outward radial force is reduced. This can be quite significant. Testing by just changing to a lower-radial-preload ring found an easy 5 hp for a four-cylinder 1,600-cc Formula Ford engine. For a V-8, making such a move can be worth double to triple this amount, depending on the RPM involved.

If you are starting on the development of a particular piston and ring combination, it is a good idea to deal with a piston ring specialist such as Total Seal because they can point you toward effecting good oil control while limiting frictional loss.

Lubrication System Although we have concluded the discussion of oil-control rings, we have not entirely finished the subject of oil control. There is plenty of opportunity to squander power in the engine’s lubrication system. All too often, the assumption is made that the oil pressure for a race engine needs to be really high. There’s an often-quoted general rule that an engine needs 10 psi for every 1,000 rpm at which it turns. Although it’s not quite as simple as that, it is at least a starting point and, for the most part, it’s at least on the safe side. But applying this basic rule to a 1,000-cc, 11,000-rpm engine means way more pressure is needed, as well as a substantial power reduction just to drive the oil pump at such pressures. Many manufacturers’ stock performance engines survive 200,000+ miles with no more than about 55 psi. With few exceptions, a typical fourcylinder, 2-liter performance engine turning up to 8,000 rpm needs only about 55 psi and maybe only 70 or so at 10,000 rpm. But oil pressure alone is not the only issue here. Be aware that a high-volume oil pump sounds better on paper than it is in practice. Ask yourself why you would want to install a higher-volume oil pump if the stock one was already into the bypass valve at 2,500 rpm. The answer here is: Maybe you shouldn’t. Increased bearing clearances, more RPM, and hotter oil may justify an increase in oil pressure by adjustments to the bypass valvespring. But that, for the most part, is the only change needed. As for bearing clearances, we find that in most instances 0.002 for the rods and 0.0025 for the main bearings work well. However, with really good parts and close manufacturing tolerances, these clearances can, as we see with Cup Car engines, be closed up by a half to one thousandth or so (0.0005 to 0.001 inch).

The DSS road-race pan being installed on a small-block Ford 5.0 engine not only has capacity for an extra quart or so of oil but is also baffled to stop the oil from becoming entrained in the rotating assembly.

Some oil scrapper/baffles can take on a quite complex form, such as seen here. This is especially so when ground clearance is an issue to take into account.

This Moroso pan has all the basic features a performance engine’s pan needs, while remaining a cost-effective part.

This Barnes dry-sump pump handles the lubrication requirements for a NASCAR Sprint Cup engine. It is comprised of one pressure and four scavenge stages. This results in a pan vacuum of about 10 inches of mercury. Almost certainly of greater importance than the search for oil pressure is what happens to the oil after it has done its job lubing parts and cooling pistons. No matter how you look at, it the bottom end of an engine is pretty ugly in terms of aerodynamics. The crank and rods can absorb quite a big

chunk of power just thrashing around in the crankcase. The number-one job here is to minimize windage and viscous losses by separating the oil from the rotating assembly. The first move is to make sure the pan is sufficiently deep so that the crank is far from being able to dip into the oil. Probably of near-equal importance is making sure that the oil pump pick-up is always well immersed in the oil. This requires using a horizontal baffle to reduce surge to a minimum. An effective wet-sump system for a serious drag-race engine has an assortment of trap doors, baffles, scrapers, and de-aerating mesh screens. The purpose here is to allow, as far as possible, the crank to run in an oil-free environment. For the most part, drag-race pans can be deep because ground clearance is not so much of an issue but, for a road race car, the situation changes somewhat.

Production pumps get the job done, but there is often enough capacity to streamline the entry and exit passages to good effect. Porting and selective assembly can account for as much as an 85-percent increase in oil pressure, up to the point the bypass opens. Even then, the pressure delivered will likely be more than stock. When ground clearance or a bigger budget permits, opting for a drysump system is well worth considering. Although far more complex in nature, a dry-sump system allows a lower installation of the engine in the chassis, thus lowering the car’s center of gravity. Dry sumping also allows us to use multiple scavenge pumps with a capacity to not only pull out all the oil but to also drop the pan/sump internal pressure well below atmospheric. Exactly what is accepted universally as the optimum absolute crankcase pressure is still far from decided. But I have seen more than 15 inches of mercury used to seemingly good effect. My own experience

indicates that anything over about 2 inches of mercury (40 inches H2O or 1 psi) starts to pay a worthwhile dividend.

10 Rules for a Successful Stroker Build 1.

Use the longest rod possible to minimize frictional losses and mechanical noise. 2. Use the lightest reciprocating components consistent with sufficient strength for the RPM. 3. Use an effective crank damper. 4. Use an oil pan that keeps the oil away from the rotating assembly. 5. Use as high a CR as possible to offset the effect of increased pistonto-bore friction. 6. Use cylinder heads with the largest valves possible. 7. Tighten up the LCA from whatever was optimum before the stroke increase. 8. Increase valve lift proportional to the stroke increase. 9. Check that the induction system has adequate flow to service the extra displacement. 10. Keep the induction system as cool as possible; this makes more difference with a stretched engine.

This test demonstrates just what can be achieved by making an oil selection based on dyno testing. This particular engine was for a NASCAR late-model stocker, where the rules called for an untouched crate motor. It was known that the hydraulic lifters used by the factory were somewhat inconsistent and a little prone to collapse. Here, I convinced Ultra Pro Machining’s oil expert to make a blend to combat lifter collapse as much as possible. The results even surprised me. The green curve is for a well-known off-the-shelf, high-quality mineral oil (402 ft-lbs and 388 hp). The blue curve is one of the best of a half-dozen synthetic race blends tested (396 ft-lbs and 399 hp). Although they consistently made more horsepower, it was suspected that they did not hold up the hydraulic lifters as well as maybe they needed to. After many attempts at manipulating the shear characteristics of the oil, UPM’s oil guru hit the jackpot. The red curve is the result (405 ft-lbs and 405 hp). The bottom line here is: For any hydraulic-cammed engine, for race or street, this is the oil to use.

CHAPTER 13

EXHAUST MANIFOLDS British engineer and author Philip H. Smith wrote a 250-page book entirely on intake and exhaust systems because the subject warranted it. We don’t have space here to do that, so you are going to get the hands-on basics of an important and complex subject. The energy available to scavenge the combustion chamber is far higher than is often supposed, and harnessing it can produce a power increase equal to 3 to 4 pounds of supercharger boost. Such a power gain is there for the taking and requires little more than getting the exhaust manifold (header) dimensions right. The principal and most critical dimensions we have to take care of are shown in Illustration 13-1. What we see here is about the simplest form of exhaust manifold: the 4-into-1 system. There are others, such as the 4-into-2into-1 or 6-into-3-into-1 systems. As you may expect, due to the number of cylinders involved and the engine’s layout, variations abound. To get a handle on how the system “tunes,” we look at a simple onecylinder system. In Illustration 13-2, we have a single exhaust pipe with the exhaust-valve end to the right and the open end to the left. When the exhaust valve initially opens, high-pressure gas from the cylinder is rapidly dumped into the pipe. The high-pressure wave then travels along the pipe at the speed of sound (about 1,300 ft/sec in hot exhaust gases).

In a high-performance engine, the exhaust has a greater effect on the induction system than the piston does going down the bore on the induction stroke. To make use of this valuable performance asset, it is necessary to understand what is going on. The fundamentals of such are dealt with in this chapter.

13-1. Dimension A (yellow line) is the primary length. Dimension B is the secondary (collector) length. Dimension C is the primary diameter and Dimension D is the secondary/collector diameter. When the wave reaches the end of the exhaust pipe, two things happen: It is reflected, and it changes from a positive-pressure wave (above atmospheric) to a negative-pressure wave (below atmospheric). This negative-pressure wave, which can be as much as 7 psi below atmospheric pressure, then travels back along the exhaust pipe to the exhaust valve. If this negative-pressure wave arrives at the exhaust valve while it is open and in the overlap period, it not only sucks out the exhaust residuals in the combustion chamber but also lowers the chamber pressure to something well below the intake-port pressure. This results in the charge flowing into the cylinder, even though the piston may not have quite reached TDC. By the time the piston is about to start the intake stroke, the intake charge can already be moving at as much as 100 ft/sec or more into the cylinder. For the events to happen at the intended RPM, the exhaust-pipes need to be a certain length. There are many formulas out there that give a good working result. These formulas take into account the valve opening and closing events as well as the RPM involved. Rather than bog you down with yet more formulas, I feel a simple graph (see Chart 13-2) gets the job done more easily. The chart is easy to use, but it does make an assumption that you need to take into account. The assumption here is that the cam being used will get

longer, but the duration of a cam and the characteristic it produces are very much intertwined with the valve-size-to-displacement ratio. For instance, a 280-degree cam in a four-valve engine looks like a small race cam, and power can peak as high as 8,000 rpm. That same 280-degree duration, in a 500-inch big-block Chevy, may peak at no more than 6,200 rpm and run like a very civilized street cam. The company that manufactured your cam should provide the RPM for peak power with the right exhaust. As for applications, this chart is for one-, two-, four-, and six-cylinder engines and in-line or flatcylinder configurations. It is not for V-8s. When it comes to primary length selection of a two-plane crank V-8, all those length calculations go out the window. What we have is actually the most complex situation, in terms of function, paired with what must be the simplest solution.

13-2. Shown here in much simplified form is the process by which the exhaust can be made to scavenge the combustion chamber and make the intake flow start much sooner.

In this graph, the intake-port velocity is computed from the measured intake pressures. Of interest here is that the intake velocity, when the piston is at TDC and the intake stroke about to commence, exceeds 80 ft/sec (nearly 60 mph). To see how this works out, it must first be understood that a two-plane crank V-8 is (unlike Cosworth’s old DFV F1 V-8) not two inline-fours put together, but two V-4s. The firing pulses down each bank do not occur at regular intervals but are, starting at cylinder number-1, spaced 0-270-18090-180-270 and so on. Chart 13-4 shows two pressure pulses seen at the entrance into the collector. Note that cylinders number-5 and -7 (being only 90 degrees apart) appear as if they are one single, much-larger cylinder. But that’s not all; that pair of cylinders also appears to the collector as if it is revving more slowly than the other pair. The result here is that a twoplane crank V-8 simply does not conform to the primary-length predictions (shown in Chart 13-3). Now that may sound a little disappointing but, by a quirk of fate, it all works in our favor. What we find in practice is that a V-8 is very much primary-length insensitive. If the primaries are somewhere between about 28 and 40 inches, things work just fine. (Fig 13-5) This is also a reason why getting all the pipes the exact same length for a V-8 is something of a waste of time. It’s best to make sure that all the bends are a big radius and the route to the collector is the least tortuous.

Pipe Diameter

The diameter chosen is equally important, and there are abundant methods that show you how to calculate it. However, none that I know of actually takes the important factor of valve lift and the flow capability of the port into account. Primary Pipe Diameter Sorting out the length needed to tune to the desired RPM really is only half the way toward achieving an effective primary pipe spec. Over the years I developed some curves explicitly based on dyno test results that appear, in most instances, to deliver very near optimal results. These curves, as seen in Chart 13-6, rely solely on the flow capability of the port at the full valve lift figure delivered by the cam.

13-3. This chart assumes a progressively longer cam as RPM across the bottom scale increases. This being so, it predicts the primary length for a short street cam at the low end of the RPM scale and a race cam for the high end. This means you must choose a realistic tuning RPM that also represents about where the cam allows peak power to occur, given a properly tuned exhaust. To make the most use of the tuning effect, choose an RPM that is about 5 percent less than the expected peak power RPM. Note: All these lengths are from the valve, so the length of the exhaust port in the head casting must be subtracted for the length computed here. Secondary Dimensions

At this point, we have sorted out the primaries. To get a workable secondary diameter, multiply the primary by 1.68 for a hot street application, 1.73 for street/strip, and 1.78 for all-out race. If we are dealing with a onecylinder engine, better results may be seen with a smaller secondary of about 1.6 times the primary. Determining the secondary/collector diameter was straightforward enough. The same applies to the collector but, because it is easy to adjust, we don’t need to take theory as the last word. For a V-8, the secondary, unlike the primary, is very sensitive to length tuning. The test in Chart 13-7 shows how just 12 inches of collector length can affect power.

13-4. Here are the exhaust pulses, as seen at the collector, from one bank of a small-block Chevy V-8. Note that, although cylinders No. 1 and No. 3 produce a distinctive pulse of their own, the pulse as seen at the collector from cylinders No. 5 and No. 7 appears to be one much bigger cylinder. This causes problems when attempts are made to tune in the same fashion as an in-line four.

13-5. From these tests on a 350 small-block Chevy, equipped with a hot street cam, we can see that primary lengths from a 29-inch average to a 38-inch average showed only minimal changes in output. The advantage of engine or chassis dyno testing exhaust systems (to optimize secondary lengths) is that adjusting the length is easy. That means there is almost no excuse for not getting on the dyno and doing just that. In the collector length tests, the blue curve was with about a 1.5-inch collector length after the 4-pipe merge section. By adding 12 inches of collector length, the output rose to that shown by the red curve. The gains are far from inconsequential: 40 ft-lbs at 3,700 rpm, and 12 hp at 6,200 rpm.

13-6. To use this chart, you need to know (or need a good estimate of) the exhaust flow at full valve lift. Using this figure, go up from the bottom of the chart to one of the curves (green for street, purple for street/strip, blue for race). An example looks like this: port flow at max valve lift on a 130cfm street/strip application. From the 130-cfm mark on the bottom scale, go up to where you intersect the purple line, and then go left to read the pipe ID—in this case 1.375.

The effect of an even moderately-too-large collector diameter can be seen from this dyno test. The largest diameter delivered less output all the way through the RPM range. Even at peak power, it was marginally less than either of the smaller-diameter collectors.

13-7. As this dyno test on a V-8 shows, secondary length is critical. On occasion, I have built V-8 exhaust systems for street use with a secondary as long as 60 inches. This proved to be capable of boosting lowspeed (2,000 to 3,000 rpm) torque as much as 20 ft-lbs over a more usual 20to 30-inch collector. At this juncture, I need to point out that the secondary length for a V-8 application is, at lower RPM, somewhat longer than for a one-, two-, or four-cylinder in-line engine. If you are working with a non-V8 application, you can get a good starting point for the secondary length by making it 15 inches less than the primary length.

Alternative Configurations So far, we have discussed basic manifold configurations based on all the primaries dumping into a secondary. This gives us the commonly seen 4into-1 system used on V-8s and high-RPM four-cylinder engines. If the engine is a four-cylinder unit, and is not likely to see the top side of 7,800 to 8,000 rpm, then a 4-into-2-into-1 system is the way to go. Such a system has the advantage of fitting the confines of the engine bay easier and produces more torque up to about 6,800 to 7,000 rpm. At the lower speeds (2,000 to 4,000), such a system can be up by as much as 12 ft-lbs per liter (0.2 ft-lbs

per cubic inch). Above 7,800 to 8,000 rpm, it’s a clear-cut advantage to go with a 4-into-1 system. The V-8 world has what can best be described as its own version of a four-cylinder 4-2-1 system. In this instance, though, it is more of a 4-2-1 collector design. As can be seen in Illustration 13-9, the system still has long primary pipes but, instead of merging into a 4-into-1 collector, each pair of primaries merges first into one short secondary and then makes another merge into the collector. As of 2010, these are the systems of choice for some of the most powerful, naturally aspirated, single-4-barrel-carbed V-8s. Just how much they are worth on an 890-hp 350, I cannot say with any certainty. But on a 740-hp 350, they showed some 10 hp more at the mid-to-top end, but maybe a little less horsepower at 500 rpm under the peak torque. For a street application, they represent a big additional dollar investment over a conventional 4-into-1, and so may not represent a cost-effective alternative.

13-8. For four-cylinder applications below 7,800 rpm, a 4-2-1 system (as seen here) gives better results than a 4-1 system. The primary and secondary lengths need to be about the same and generally run around the 13- to 18-inch mark.

Using This Header Tech Because of space limitations here, I have taken some short cuts concerning exhaust manifolds. However, most hot rodders do not make their own manifolds and, instead, buy them either as off-the-shelf components or as custom-built items. This being the case, the intent of this chapter was more to illustrate what goes into the selection of a manifold. This chapter should leave you, as a prospective customer, with a good idea of the implications involved if you make a poor choice. Heed what I say here and listen to the advice of someone who has expertise with exhaust systems for the particular engine type you are working with. Above all, given the choices, I advise you err on the conservative side. The one common theme all through the dyno tests shown in this chapter is that, past a certain point, bigger is not better—a little too small does not hurt to any measurable extent.

13-9. This four-long-tube V-8 manifold (header) utilizes a 4-2-1 merge collector. Carefully conducted dyno testing indicates about a 1.2-percent increase in output above the peak torque RPM.

This LS6 Chevy’s power output climbed from 739 hp with the original 4-1 collector to 750 hp with the change to the 4-2-1 merge collector shown here.

CHAPTER 14

MUFFLERS TO TAIL PIPES

Building an exhaust system along the guidelines shown in this chapter proved an effective method of muffling a very high output engine down to acceptable street levels without losing any power.

In 1980, I built a 400-ft-lb, 404-hp 350 replacement for the anemic 158hp 305 in my California Pontiac Trans Am. Back then, it took real effort to exceed 400 ft-lbs and similar horsepower. Imagine my disappointment when, no matter what I used for mufflers, the output dropped by some 20 ft-lbs and 25 hp. But my experience as a co-designer of a no-loss system for the original Mini Coopers encouraged me to try the same stunt for V-8application mufflers. Aided by an acoustics expert, the result was the Sonic Turbo manufactured by Cyclone (a division of Walker/Dynomax). A big Hot Rod magazine muffler shootout at Gale Bank’s facility demonstrated that my 2¼-inch Sonic Turbo’s muffler outpaced everybody else’s 2½-inch mufflers. This sparked a sales run of hundreds of thousands, which seemed to inspire the industry into more aggressive muffler design efforts. Now, thirty years later, you, as a hot rodder, inherit that legacy. These days, all the parts to build a no-loss system are at hand and affordable. What seems so often to be lacking is system tech know-how; but here, that’s about to be set right.

Simple Steps to Success We dealt with the exhaust manifold (headers) and collector system’s role in power production in Chapter 13. From what was learned there, it should only be a short step to appreciating how parts selection and positioning of resonators, routing pipes, crossovers, mufflers, and the like will be a winning factor. This will be especially so if catalytic converters are involved in the equation.

In this chapter, we deal with catalytic converters (green-tipped arrows), Xpipes and crossover pipes (black-tipped arrow), entry and exit pipes (redtipped arrows), and mufflers (blue-tipped arrows).

For an emissions-legal street exhaust system, catalytic converter flow is the biggest problem. Be sure to diligently research the flow capability of whatever you buy. I initially wrote on no-loss exhaust systems as early as 1980, so I am surprised that, as of 2010, it is still commonly believed that high output and minimal noise are mutually exclusive. As detailed here, a quiet system that allows within 1 percent of open exhaust power is practical, and can easily deliver a 40-plus-hp advantage over your less-informed competition.

Cats and Mufflers If you are attempting to build an exhaust system that has to be equipped with cats, then you are in for a rough ride. But it is doable, so let’s start at square one. Essentially, there are two key factors to avoiding exhaustsystem-induced power loss. Inappropriate catalytic converter and/or muffler selection and installation effectively cancels any advantages system length/diameter tuning could have delivered. The questions most often asked are: What does it take to get it right? How much power are we likely to lose if the system is optimal? The simple answers are: “Not much” and “Zero.” To build a no-loss muffled, high-performance/race system, it is vital we work with the two key exhaust system factors in total isolation from each other. The first of these factors is the pressure-wave tuning from our length/diameter selection, and the second is minimizing backpressure by selecting cats and mufflers with suitable flow capacity for the power concerned. For the most part, this involves nothing more than knowledgeable component selection and installation. Cat and Muffler Flow Carbs (at least in the United States) are selected based on flow capacity rather than size, simply because the engine does not know size-only flow. This begs the question as to why the same should not apply to muffler selection. The answer (muffler manufacturers: please note) is that it should, as the engine’s output is influenced minimally by size but dramatically by flow capability. A muffler purchase based on pipe diameter has no

performance merit. The only reason for knowing the muffler pipe size is for fitment reasons.

Evaluating muffler flow, and modifying as required to optimize it, pays dividends in output. As unlikely as it may seem, not all muffler companies do this. The engine cares nothing about muffler pipe size but it really cares what the muffler flows, and that is dictated by its internal design. From this, it follows that the informed hot rodder/engine builder should select cats and mufflers based on flow, not pipe size. In other words, positive results are based on flow, not on pipe diameter. There are many ways to enhance the accomplishment of sufficient flow to get the job done. We can firmly establish that its flow, not its size, dictates the success of the exhaust system. In Illustration 14-1 we see a clear demonstration of how an ineffective component, such as the cat or the muffler, can influence the apparent pipe size seen by the engine. Let’s start by viewing a muffler installation as three distinct parts. In the top left drawing are the in-going pipe, muffler core, and exit pipe. Assembled, they appear as in the top right drawing. In the middle left drawing is a typical muffler. Due to a design process apparently unaided by a flow bench, the muffler’s core flow is significantly less than an equivalent length of pipe the size of the entry and exit pipe. Because the core flow is less than the entry and exit pipe, the engine sees the muffler as if it were a smaller (and consequently more restrictive) pipe like the reduced-diameter tube at the middle right.

X-pipes and crossover tubes allow a measure of muffler flow-capacity sharing. This helps reduce backpressure and improves noise reduction. The “core” flow of this muffler is the prime restriction and has little, if anything, to do with the inlet and outlet diameter of the muffler itself. What we need for best performance is a muffler (or cat, for that matter) that simulates the lower pipe where the core flow is actually greater than the inlet and exit pipe. Result: The muffler is seen by the engine as a near-zero restriction. A section of straight pipe the length of a typical muffler, rated at the same test pressure as a carb (1½ inches of mercury), flows about 115 cfm per square inch. Given this flow rating, we see about 560 cfm from a 2½inch pipe. If we have a 2½–inch muffler that flows 400 cfm, the engine reacts to this just the same as it does to a piece of straight pipe flowing 400 cfm. At 115 cfm per square inch, that’s the equivalent of a pipe only 2.1 inches in diameter. This apparent pipe size concept is important to appreciate. The reason is that so many racers fixate on having as large a pipe into and out of the muffler as possible. As you should now see, this concern is totally misplaced; in all but a few cases, the muffler is the point of restriction, not the pipe. The fact that muffler core flow is normally lower than the connecting pipe can be offset by installing something higher flowing, such as a 4-inch muffler, into an otherwise 2¾-inch system.

Here is a dyno test of the Magnaflow C6 Corvette system. Basically, we have a 7 to 11 ft-lbs increase in output occurring over the entire RPM range. Peak power climbed by 11 hp.

14-1. For a clear explanation of what is being shown here, refer to the text.

Even as inexpensive as it is, this Walker Dynomax muffler is very effective. It provides good flow and output while suppressing noise. The design is a prime example of the extensive use of flow-bench results.

This stainless-steel-construction Walker straight-through muffler may not look the most glamorous you’ve ever seen, but high flow and good noise suppression results in relatively quiet power.

Muffler Flow: How Much is Needed? The first point to appreciate here is that optimally sized collectors/secondary pipes are not sized so as to meet the engine’s flow requirement, but more by the need to produce the desired pressure-wave characteristics. For instance, a 700-hp V-8 engine may have a dyno-

optimized 3¾-inch-diameter collector. This diameter, in conjunction with the length, resulted in the system tuning-in at the desired RPM. But from the standpoint of flow, a 3-inch pipe from each bank is capable of handling all of such an engine’s flow requirements. Without data to the contrary, it seems safe to assume that the more a muffler flows, the better. This, fortunately, is not so and here’s why: Increasing muffler flow unlocks potential engine power. Once all the potential power is unlocked, further increases in exhaust system flow do not produce any further benefits in terms of power. But what may be good for power may not be good for noise; any excess flow capability can lead to a noisier system. From this, we can conclude that too much muffler flow serves no useful purpose and possibly costs more money than necessary. The trick here is to use just the right amount of muffler—no more and certainly no less. This allows the full power potential of the engine to be realized at the lowest cost, without undue compromises in terms of noise. Now the question is: How much flow is enough? In about 1985, many race-series officials anticipated that mufflers would be mandated for race vehicles, so I embarked on a series of tests to determine a race engine’s minimum-flow threshold. Performing such tests might look easy, but to get meaningful results, it is very important to separate, as far as possible, the effects of flow and backpressure from the effects of pressure-wave tuning. It is entirely practical to do this by means of a pressure-wave termination chamber, which for all practical purposes is a very-oversized resonator box. Knowing when and how to use a big resonator box can be a great aid in building a high-performance system.

14-2. This chart shows the effect that flow and backpressure have on the power output when their effects are separated from the pressure-wave tuning due to header and collector sizing. These numbers were developed from several engines with cams ranging from about 290- to about 298degrees duration. As can be seen, after the flow exceeded about 2.2 cfm, all the power from increased flow had been unlocked; so added flow produced no further increases.

This 3-inch Borla muffler flowed the same as a straight piece of pipe of equal length because its core flow was higher than the inlet- and exit-pipe flow. To get an idea of how lack of sufficient flow, and consequently backpressure, can affect output, check out the results of Chart 14-2. The curves shown are the results of tests run on a number of engines of various types. The only common element of significance between these engines was the use of a cam of 290 degrees or more of seat (advertised) duration and CRs of 10.5:1 or more. From the curves on this graph, you can see that the trend is: As flow is increased (and consequently backpressure reduced) on an initially flow-restricted engine, power increases rapidly at first and then gains diminish. Once the available flow exceeds about 2.2 cfm per hp, increasing muffler capacity drops the possible gains to less than 1 percent. Having established, within reason, from the graph that 2.2 cfm per openpipe horsepower means zero loss from backpressure, we can determine how much muffler flow a particular engine needs based solely on its open-pipe horsepower. Just make a reasonable estimate of the engine’s open exhaust pipe power potential and multiply by 2.2. For example, a V-8 making 600 hp on open exhaust will require 600 × 2.2 = 1,320 cfm. Two 660-cfm mufflers will get the job done and contain the backpressure-induced power loss to 5 hp or less. From the foregoing, with mufflers rated in CFM, you can see how easy making an appropriate choice gets. Before leaving the subject of muffler flow requirements, I advise against letting anyone convince you that a muffler system has to have some backpressure for the best output. To put it bluntly, that’s just a bunch of baloney.

Pressure Waves Now, we can move on to methods critical to utilizing suitable-capacity mufflers in the system without disrupting the length-induced pressure-wave tuning. Probably the best way to ease into this somewhat complex subject is to consider some of the published muffler test results done between about 1990 and 2008. At face value, these tests indicated that sometimes lower flow mufflers induced at least some backpressure, and that resulted in the test engine making its best power. In all such tests that I studied, the

conclusion (as opposed to the tests) were incorrect. There are a number of reasons for this and all are relevant to building a no-loss exhaust system.

14-3. From this illustration, it can be seen that simply adding a straightthrough glass-pack muffler to a length-tuned system disrupts the optimized secondary length. This being the case, no matter how well the glass pack may flow, power is reduced. Conversely, the open internal design of a Flowmaster-style muffler allows the effect of the tuned length to remain unchanged. The first point cancels the supposed validity of these back-to-back test results because of (see Illustration 14-3) the various internal muffler designs involved. Many mufflers consist of a number of inter-connected chambers, each presenting the exhaust a different ease of access. Others are the glasspack variety. These types represent opposite ends of a spectrum and have a substantially differing response to arriving pressure waves. When we discussed collector lengths in Chapter 13, I emphasized that, in most cases, it was more critical than the primary pipe lengths. Adding even a zero-backpressure muffler to a system with already-optimized lengths can alter the pressure-wave response. The added length can simply tune the exhaust out of phase with your requirements and result in a drop in power.

The technique to use here is to install mufflers that don’t alter the tuned lengths of the system. Let us assume the mufflers being tested are attached directly to the end of the collector. A pressure wave is reflected either at the end of the exhaust pipe or when a sizable increase in cross-sectional area occurs. Open-chamber mufflers, such as Flowmasters, often appear to the pressure wave much the same as the end of the pipe. The result is that the pressure waves see no change in length, and reflection occurs much as it did prior to installing the muffler.

This Magnaflow muffler is a straight-through type. It has extremely good flow characteristics but you need to preserve the collector-tuned lengths. Because of its intrinsically differing design, a glass-pack muffler acts significantly different. As the exhaust gases enter, they do not see what appears to be the end of a pipe so much as an extension of the collector. The result is a reduction of power, even though there is no measurable backpressure involved. From this, we can safely conclude that most comparative muffler tests were, in fact, “pseudo pipe length” tests. Although many invalid conclusions were drawn, these tests still demonstrated some important facts. The most important is that the engine’s needs, in terms of flow and pressure-wave-length tuning, must be isolated, one from the other. This is easy to do by means of a pressure wave termination box (PWTB, or resonator box). Incorporating a big resonator box into a system produces a layout along the lines seen in Illustration 14-4. Given adequate volume, the PTWB makes everything downstream appear invisible to the header’s primary and secondary tuned lengths. With a flow capability of 2.2 cfm or more, the

muffler also appears virtually invisible from the flow standpoint. As a result, we have a muffled system that produces virtually the same power as an open exhaust.

Ideally, the PWTB needs to be shaped as in the upper drawing. If it is being used for a V-8 engine configuration, then it is advantageous to use a balance tube, preferably at position No. 1. Failing that, at No. 2, but is not usually quite as effective. An empty muffler (No. 3) can be used, but the exit pipe must have a suitable lead in, or flow losses result. The point of all the muffler selection and installation tech discussed so far is to produce an acceptably quiet system. If it does not do that, the point of the exercise is lost. By using no more muffler flow than needed, we are giving whatever mufflers are selected the best chance of doing the job. Unfortunately, from one engine type to another, mufflers can be a little inconsistent and unpredictable in terms of noise suppression. Engines with high compression ratios and long-duration cams are usually more demanding in terms of noise reduction. Big cubes, shorter cams, and lower CRs are easier to muffle. There will inevitably be a system built to all the right specs that fails to reduce noise to the levels hoped for. A good start here, in order to avoid this as much as possible, is to peruse some of the bigger muffler companies’ web sites. On these sites you can see and hear chassis dyno tests

of a wide variety of mufflers (including stock) on an extensive range of vehicles. Also be aware that system installation can also affect the sound level experienced, especially in the vehicle’s interior. Tail pipes ending under the car’s bodywork use it as a sound box in much the same way as the strings cause a guitar body to resonate. Your best option is to have the tail pipe(s) go all the way to the rear or have a side exit. If the rear exit is used, further noise reduction, however small, can be had by downturned exit pipes, angled slightly in toward each other. As far as output goes, the post-muffler tail pipe length has no measurable effect on the power if the large change in cross section discussed previously is present upstream (toward the motor) of the tail pipe. An open-type muffler or a resonator box provides this cross-sectional change. The tail pipe length exiting most glass-pack installations is also of little consequence if a resonator box is used, but can be of significant influence if not.

14-4. Here is a move I found worked when I had the wrong-size Flowmaster for the tail pipes involved, but had some adaptors on hand to do the install anyway. The result was a small but measurable increase in low-speed torque with no downside at the top end. I presumed this was from the anti-reversionary effect of the smaller collector pipe entering an intrinsically larger muffler. Power seen at top end was still the same as that on open pipes.

Building a system with these guidelines proved to be an effective method of muffling a very-high-output engine down to acceptable street levels without losing any power.

Crossover and Balance Pipes The addition of a balance or × pipe can refine nearly all V-8 exhaust systems. These have two potential attributes: reduced noise and increased power. Extensive dyno testing of both these factors has indicated balance and × pipes are 100-percent successful at reducing noise. The reductions measure from a minimum of 1 dB to a maximum of 3 dB, with 2 dB being common. As far as power is concerned, things are a little less certain. With engines between about 325 and 550 hp, experience indicates that, in about 60 percent of the cases (mostly with balance pipes), the engine can deliver as much as 12 additional hp with 5 to 8 being the most common. The other remaining 40 percent tested showed virtually no change in output, either up or down. Based on such results, we can conclude that a balance or × pipe is always a positive asset and never a negative. Balance-pipe sizing seems to be not overly critical. The only really influential dimension is the pipe diameter. This needs to have an area at least equal to that of a 2¼-inch-diameter pipe (4 in 2) with 2½ to 2¾ being preferable. Though limited to tests on engines up to a little less than 600 hp,

there seems to be no measurable benefit to using a crossover pipe bigger than 2¾ inches in diameter. As for the crossover length, dyno results indicate that 18 inches long responds in virtually the same manner as 72 inches long.

The Ultimate System? The system above is one I designed for a 700-hp, naturally aspirated street/strip small-block Chevy installed in a 1986 Corvette. It produced acceptable street-noise levels without any measurable drop in power. Although you may have to adopt some slightly different steps toward getting an acceptable installation, keeping sight of the principles involved delivers similar results. Step outside the guidelines, and you are on your own.