Engine Dyno Guide

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Engine Builder’Engine Builder’s Guide s Guide TTooFilling & Emptying SimFilling & Emptying SimulationulationAnd High-PAnd High-Perferformance Designormance Design

MINI-GUIDEMINI-GUIDE

DYNOSDYNOSCOMPLIMENTARYYDYNOSDeskTopDeskTopDeskTop

®

Motion Software, Inc.535 West Lambert RoadBuilding “E”Brea, California 92821

ISBN 1-888155-00-0PART No. M43

INCLUDES:

Details Of All On-Screen Menu Choices/Selections

Motion’s Filling & Emptying Simulation Basics

Cylinderhead Flow And Discharge Coefficients

Understanding And Calculating Camshaft Timing

Ram Tuning And Pressure Wave Dynamics

Expose Fallacies About Engine Components

Complete Glossary And Much, Much More!

2—DeskTop Dynos Mini Guide

PLEASE READ THIS LICENSE CAREFULLY BEFOREBREAKING THE SEAL ON THE DISKETTE ENVELOPE ANDUSING THE SOFTWARE. BY BREAKING THE SEAL ONTHE DISKETTE ENVELOPE, YOU ARE AGREEING TO BEBOUND BY THE TERMS OF THIS LICENSE. IF YOU DONOT AGREE TO THE TERMS OF THIS LICENSE,PROMPTLY RETURN THE SOFTWARE PACKAGE, COM-PLETE, WITH THE SEAL ON THE DISKETTE ENVELOPEUNBROKEN, TO THE PLACE WHERE YOU OBTAINED ITAND YOUR MONEY WILL BE REFUNDED. IF THE PLACEOF PURCHASE WILL NOT REFUND YOUR MONEY, RE-TURN THE ENTIRE UNUSED SOFTWARE PACKAGE,ALONG WITH YOUR PURCHASE RECEIPT, TO MOTIONSOFTWARE, INC. AT THE ADDRESS AT THE END OF THISAGREEMENT. MOTION SOFTWARE, INC. WILL REFUNDYOUR PURCHASE PRICE WITHIN 60 DAYS. NO REFUNDSWILL BE GIVEN IF THE DISKETTE ENVELOPE HAS BEENOPENED.

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Motion Software, Inc.535 West Lambert, Bldg. EBrea, CA 92821714-255-2931; Fax 714-255-7956

© 1995, 1996 By Motion Software, Inc. All rights reserved. Motion Soft-ware, Inc., MS-DOS, DOS, Windows, and Windows95 are trademarks ofMicrosoft Corporation. IBM is a trademark of the International BusinessMachines Corp. DeskTop Dyno and DeskTop Dynos is a trademark of Mo-tion Software, Inc.

Please report all copyright violations to Motion Software, Inc. at the aboveaddress and to:Software Publishers Association1730 M Street, N.W.Suite 700, Washington, DC 20036-4510800-388-7478

M

DeskTop Dynos Mini Guide—3


DYNOSDYNOSDYNOSDeskTopDeskTopDeskTop

COMPLIMENTARYY

BY:

LARRY ATHERTON


Copyright © 1996 by Motion Software, Inc.,535 West Lambert, Bldg. E, Brea, CA 92821.All rights reserved. All text and photographsin this publication are the copyright prop-erty of Motion Software, Inc. It is unlawfulto reproduce—or copy in any way—resell,or redistribute this information without theexpressed written permission of MotionSoftware, Inc. Printed in U.S.A.

The text, photographs,drawings, and otherartwork (hereafter referredto as information) con-tained in this publication issold without any warrantyas to its usability orperformance. In all cases,original manufacturer’srecommendations, pro-cedures, and instructionssupersede and takeprecedence over descrip-tions herein. Specificcomponent design andmechanical procedures––and the qualifications ofindividual readers––are

beyond the control of the publisher, therefore the publisher disclaims all liability,either expressed or implied, for use of the information in this publication. All risk forits use is entirely assumed by the purchaser/user. In no event shall Motion Software,Inc. be liable for any indirect, special, or consequential damages, including but notlimited to personal injury or any other damages, arising out of the use or misuse ofany information in this publication.This book in an independent publication. All trademarks are the registered propertyof the trademark holders.The publisher (Motion Software, Inc.) reserves the right to revise this publication orchange its content from time to time without obligation to notify any persons of suchrevisions or changes.

Motion Software, Inc., 535 West Lambert, Bldg. E, Brea, CA 92821

ISBN 1-888155-00-0PART No. M43

TECHNICAL CONSULTANT:

CURTIS LEAVERTON

WRITTEN AND PRODUCED BY:

LARRY ATHERTON

TECHNICAL DRAWINGS BY:

JIM DENNEWILL


DYNOSDYNOSDYNOSDeskTopDeskTopDeskTop

COMPLIMENTARYY

Engine Builder’sGuide To

Filling & EmptyingSimulation

®


SOFTWARE LICENSE .................................................. Inside Front Cover

INTRODUCTION.......................................................................................... 6

USER'S GUIDE TO THE FILLING-AND-EMPTYING SIMULATION ......... 8FILLING-AND-EMPTYING BY ANY OTHER NAME ........................... 8MOTION ENGINE SIMULATION BASICS ........................................... 9THE ON-SCREEN MENU CHOICES ................................................. 10THE BORE/STROKE MENU .............................................................. 11WHAT’S A SHORTBLOCK ................................................................ 11BORE, STROKE, & COMPRESSION RATIO ................................... 11BORE AND STROKE VS. FRICTION ................................................ 12FALLACY ONE: STROKE VS. PUMPING WORK ............................ 14FALLACY TWO: LONG VS. SHORT STROKE ................................ 18CYLINDERHEAD AND VALVE DIAMETER MENUS ....................... 20CYLINDERHEADS AND DISCHARGE COEFFICIENTS .................. 21RAM TUNING AND PRESSURE WAVES ......................................... 24SORTING OUT CYLINDERHEAD MENU CHOICES ........................ 28WHEN TO CHOOSE SMALLBLOCK OR BIG-BLOCK HEADS ...... 31VALVE DIAMETERS AND AUTO CALCULATE .............................. 33THE COMPRESSION RATIO MENU ................................................. 34COMPRESSION RATIO BASICS ...................................................... 35CHANGING COMPRESSION RATIO ................................................ 36WHY HIGHER C/R PRODUCES MORE POWER ............................. 37OTHER EFFECTS OF INCREASING C/R ......................................... 38COMPRESSION RATIO ASSUMPTIONS.......................................... 38THE INDUCTION MENU .................................................................... 39AIRFLOW SELECTION ...................................................................... 39AIRFLOW MENU ASSUMPTIONS .................................................... 41INDUCTION MANIFOLD BASICS...................................................... 42MANIFOLD SELECTION ADVICE ..................................................... 44THE EXHAUST MENU ....................................................................... 52WAVE DYNAMICS IN THE EXHAUST SYSTEM .............................. 53EXHAUST MENU SELECTIONS ....................................................... 56THE CAMSHAFT MENU .................................................................... 61CAM BASICS ...................................................................................... 62VISUALIZING & CALCULATING VALVE EVENTS .......................... 63HOW VALVE EVENTS AFFECT POWER ......................................... 68CAMSHAFT MENU—LIFTER CHOICES ........................................... 75CAMSHAFT MENU—SPECIFIC CAMSHAFTS................................. 78ENTERING 0.050-INCH & SEAT-TO-SEAT TIMING ........................ 81CAMSHAFT ADVANCE AND RETARD ............................................ 84CALCULATING VALVE EVENTS ...................................................... 86

APPENDIX—A: COMMON QUESTIONS AND ANSWERS .................... 88SOFTWARE SUPPORT FAX PAGE.................................................. 97

APPENDIX—B: GLOSSARY .................................................................... 98APPENDIX—C: BIBLIOGRAPHY........................................................... 108OTHER MOTION SOFTWARE PRODUCTS .......................................... 110

CONTENTS


INTRODUCTIONE n g i n e B u i l d e r ' s G u i d e T o S i m u l a t i o n & D e s i g n

™

his is not a typical “hot rodding”book. Yet it is written specifi-cally for you: The amateur orprofessional automotive enthu-

ing engine simulations is simply anotherapproach to understanding IC engines.

But engine simulations have beenaround for years, and there are manytalented engineers, developers, and uni-versity professors that understand thephysics and math that make the IC en-gine tick. What’s very rare is the indi-vidual who has this knowledge and un-derstands racing engines and the needsof the high-performance engine builder.That individual is Curtis Leaverton, thedeveloper of Motion Software’s enginesimulation and my associate in the de-velopment of these books. This MiniGuide is, to a great extent, a transcrip-tion of a small portion of his vast knowl-edge, with my contribution being one oforganizer and presenter.

Despite the inauspicious beginningsand its odd evolution, these books havebecome true “hot rodders” engine guides.Engine books—the better ones, atleast—are a compilation of the learnedexperiences of the author/racer. Theydescribe what components and proce-dures are known to work, and they con-jecture about what is less understood orwhat seems to defy logic. This “outside-in” approach is completely understand-able. It is a direct result of the mostcommon method of engine developmentused by engine builders and racers: Trial-and-error. The DeskTop Dynos booksfollow a new path. They describe enginefunction from the “inside-out.” You won’tfind a list of bolt-on parts that work ononly a few engines, rather you’ll discoverinformation that will help you understand

Tsiast interested in both engines and com-puter simulations. This book, and it’s “big-brother” the complete DeskTop Dynosbook, may reveal more to you about high-performance and racing engines thanyou’ve discovered in all the other “en-thusiast” books and magazines you’veread. I know this to be true, because Ihave been an engine and book enthusi-ast all my life, and the nearly two-yearprocess of writing these books taught memore about high-performance enginesthan my thirty years of engine building,tinkering, and racing.

Most books, like many products aimedat the consumer, begin with a publisher(a manufacturer) “discovering” a need inthe marketplace. Every once in a while,though, a book is published that falls out-side of the typical form. It is born from aunique collaboration—the result of fortu-itous events—that could never have beenplanned. This is one of those books.

Two years ago, the DeskTop Dynosbook and Mini Guide were simply to bea comprehensive guide for Motion Soft-ware users of Filling-And-Emptying en-gine simulation software. What the au-thor failed to realize at first was that thereis very little difference between under-standing engine simulations and under-standing engines themselves. In retro-spect, it’s obvious: If a simulation is agood one, it must mimic the physics atwork inside the IC engine. Understand-


WHY engines “need” parts of a specificdesign to optimize horsepower.

Admittedly, these books only “scratchthe surface” of the vast and complexfields of engine-pressure analysis, ther-modynamics, and gas dynamics, butthat’s actually what makes these books“work.” There are many other publica-tions that delve deeply into these sub-jects, the Bibliography on page 108 lists

This book simply would not havebeen possible without the technical as-sistance of Curtis Leaverton. Over the lastfew years I have spent many enjoyabledays with Curtis discussing engines andsimulation science. He is one of only ahandful of people I’ve encountered in mylife who can explain complicated subjectswith such enthusiasm that they not onlybecome clear to the listener, but the sheerjoy of being guided toward the under-standing becomes an indelible memory!In short, Curtis is a terrific teacher (and Ilike to think, my friend). Occasionally, heconducts seminars throughout the coun-

Acknowledgments

try, and I encourage every reader of this book to avail themselves of the unique opportunity tobe “taught by one of the masters.”

I also wish to thank the many manufacturers that provided photos and information for thisbook, especially Harold Bettes of SuperFlow Corporation. Harold, a lot like Curtis Leaverton,has no shortage of enthusiasm and excitement about engines and performance. He has alsobeen extremely generous with his valuable time, and I have learned a great deal from themany hours of conversation we have had (over some great Chinese dinners!). Thanks Harold.

A heartfelt thanks to all the manufacturers that helped with this publication:Borla Performance, Bonnie SadkinCarroll Supercharging, Sheryl DavisChilds & Albert, Raymond AkerlyEdelbrock, Tom DuferHKS Performance, Howard LimHooker Headers, Jason Bruce

Intel Corporation, Rachel StewartMoroso Performance, Barbara MillerSnap-on Tools, David Heide, Tami ValeriSuperFlow Corporation, Harold BettesTPS, Bob HallWeiand, Jim Davis

Another special thanks to Jim Dennewill for the many fine drawings throughout this book.Jim’s talents and patience are both rare and boundless. Thank you for all your help and yourfriendship.

Finally, a sincere thanks to Paul Hammer. I can’t imagine what this book project wouldhave been like without Paul’s dedicated help in “putting out the fires” around me (and I canimagine some pretty awful stuff!).

a few of these tomes, and many are notespecially “approachable.”

Hopefully, what you find here will bethe right mix of theory and practical ap-plication. If you are able to gain a deeperunderstanding of the IC engine fromthese pages and from using the Desk-Top Dyno software, I will consider thisproject a success.

Larry Atherton


Filling & EmptyingBy Any Other Name

E n g i n e B u i l d e r ' s G u i d e T o S i m u l a t i o n & D e s i g n

™

USER'S GUIDE

Motion Software recently releasedengine simulation software forIBM-compatible “PC” computers.These simulation programs are

based, as of the writing of this book, onthe Filling-And-Emptying method of full-cycle analysis as applied to the 4-strokeIC (internal combustion) engine. These soft-ware packages are available under vari-ous names, including the DeskTop Dyno(available from Mr. Gasket Corporation andfrom Motion Software, Inc. and its distribu-tors) the GM PC Dyno Simulator (availablefrom GM Performance Parts) and others.Each of these packages has individualfunctional differences, but all of these prod-ucts use the same simulation methodologyand many of the menu choices are com-parable. Because of these similarities, theinformation in this guidebook appliesequally to version 2.0, 2.5, and later simu-lation software releases. Minor functionaldifferences will be discussed as required.Most screen illustrations in this book de-

MotionSoftware,Inc., hasrecentlyreleasedenginesimulationsoftware forIBM-compat-ible “PC”computers.Theseprograms are

pict the latest version of the software (ver-sion 2.5.7—in the final stages of develop-ment when this book was published). Thisguide is intended to help you obtain aclearer understanding of overall programfunction, what is covered by each menugroup, the implications of individual choiceswithin the menus, and how to interpretsimulation results. In addition, we will illus-trate what can and cannot be modeled andwhy, the drawbacks of the current simula-

based on theFilling-And-Emptying methodof full-cycleanalysis. They areavailable undervarious names,including theDeskTop Dyno , theGM PC DynoSimulator andothers.


tion software, and what advances may bepossible in the near future.

Motion Engine Simulation Basics

Motion engine simulations have beenoptimized for rapid “what if” testing. En-gine component parts are grouped inmenus along the top of the screen. One ormore selections from each menu groupeffectively “builds” a test engine (called a“paper engine” in simulation parlance).Then, simply pressing “RUN” begins a full-cycle analysis of cylinder pressures. Thesimulation plots a graphic, on-screen rep-resentation of horsepower and torque foreasy visual analysis. It is a simple proce-dure to save the test, select a new part orcomponent dimension from one of themenus, and rerun the simulation to per-form back-to-back testing. The changes inpower and torque that would occur if a“real” engine were built and tested with thecomponents at hand are clearly displayedin the graphic curves. Remarkably, thisentire process often takes less than oneminute on most computer systems.

Ease of use has always been an impor-tant design element. Without it, many ofthe tens of thousands of automotive en-thusiasts that have successfully “assembledand tested” engines would not have beenable to use our simulation programs. Imag-ine if one of the necessary inputs re-quested: “Enter the intake port flow at each0.010-inch of the valve lift up to the maxi-

Motion’s Filling AndEmptying simulationoffers an inexpensive andrapid way predict powerand torque to within 5%of actual dyno figures.The various choices inthe Bore/Stroke menu areshown here (version2.5.7—in developmentwhen this guidebook waspublished—includesexpanded menu choicesand several other en-hancements, including aCam Math Calculator ).

mum valve lift.” Yet this information isabsolutely necessary to perform a true en-gine simulation. To make this softwaretechnology available to as wide an audi-ence as possible, the information requestedby the program is straightforward and gen-erally known by most enthusiasts. Theprogram uses this “basic” information toderive flow curves, cam profiles, frictionalmodels, induction characteristics, and other“technical details.” While the widely under-stood terms in the menus have made theseprograms accessible, their less-than-exactwording has left some experienced usersscratching their heads. Engine builders andother advanced users are often very awareof the “internal complexities” of the ICengine and may not realize at first glancehow Motion’s simulations handle many ofthese important considerations.

Furthermore, some performance enthu-siasts are disappointed when they realizethat engine simulations do not include someof the components they were expecting totest. A short list of these might include:block and head metallurgy, piston typesand dome shapes, head-gasket thickness,combustion-chamber shapes, oils, oil-pandesigns, ignition timing, bolt torque loads,and many more. Modeling many of theseelements would require very complex tech-niques (with concomitant complex inputsfrom the user, not to mention extend cal-culation times) and would reveal only rela-tively small power differences. These limi-tations are discussed throughout this guide,


but a good example of complexity vs. prac-ticality can be found in oil pan testing. Inorder to simulate the conditions inside anoil pan, here are a few of the inputs thatthe user would have to address: dimen-sions of pan and lower crank/block con-tours, position of oil pump, positions andsizes of baffles, trays and screens, oil vis-cosity and temperature, level of oil in pan,acceleration and directional vectors (andhow these vectors change over time), andmore. Simply gathering together theneeded information would be quite aproject!

While it would be wonderful if an inex-pensive computer program could simplyand quickly zero-in on the optimum combi-nation of all components for any intendedapplication, that time has not yet come.However, most of the engine componentsthat play a major role in power productionare modeled in Motion’s Filling-And-Emp-tying simulations. Cam timing, compressionratio, valve size, cylinderhead configura-tion, bore and stroke, induction flow, andmanifold type are just some of the mod-eled elements that have major effects onengine power.

The simulation discussed in this guideis easy to use and provides a remarkablelevel of predictive power. The designershave purposely avoided complex areas thatwould either make data entry difficult orgreatly extend computational times. If youhave a tendency to “dismiss” these simu-lations because they seem to consist only

The simulation de-scribed in this guide is

easy to use and pro-vides a remarkable level

of predictive power. Ifyou have a tendency to

“dismiss” this simula-tion because it seems to

consist only of simplepull-down menu choices,

we encourage you totake a second look. Webelieve this software, atunder $50, offers terrific

value!

of simple pull-down menu choices, weencourage you to take a second look.Motion’s Filling And Emptying simulationoffers an inexpensive and rapid way toselect component combinations that pro-duce power and torque curves often within5% of optimum for applications that liewithin the range of the simulation. We be-lieve that’s quite an accomplishment for asoftware program you can load in your PCfor under $50!

That brings us to the main purpose ofthis guidebook. The information presentedhere revolves around detailed descriptionsof every item listed in the on-screen com-ponent menus. You’ll discover the assump-tions made by Motion programmers withrespect to each of the possible choices andcombinations. Within sections that discusseach menu category, you’ll also find sub-stantial background information that can behelpful for both your simulated and real-world engine building projects. This infor-mation was compiled from the feedback ofthousands of users, hundreds of betatesters, and countless hours of testing andexploration. We are confident that what youfind here will make using Motion enginesimulation software easier and your engineanalysis more productive.

THE ON-SCREEN MENU CHOICES:

After starting Motion’s engine simulation,the user can follow two paths of enginetesting: To recall a previous test (using the


UTILITY menu) or to “build” an engine fromscratch. Assuming the latter selection, acommon choice is to start with the leftmostBore/Stroke menu then work, menu-by-menu, from left to right. While there are norestrictions dictating the order in whichmenus must be opened, the upcomingsections in this guide follow the “natural”left-to-right progression taken by most “en-gine builders.”

THE BORE/STROKE MENU

The Bore/Stroke menu is located onthe left end of the menu bar. By openingthis menu, you are presented with a vari-ety of “pre-loaded” engine configurations.If any one of these choices is selected, theappropriate bore, stroke, and number ofcylinders will be loaded in theSHORTBLOCK section of the on-screenComponent Selection Box. In addition toselecting an existing engine configuration,you can scroll to the bottom of the Bore/Stroke menu and choose the “Other...”option. This closes the menu and positionsthe cursor in the Component Selection Box,permitting direct entry of bore, stroke (ininches to three decimal places) and num-ber of cylinders. At each data input posi-tion, a range of acceptable values is dis-played at the bottom of the screen. As withall numerical input, only values within therange limits will be accepted by the pro-gram.

What’s A “Shortblock”

When a particular engine combinationis selected from the Bore/Stroke menu, thebore, stroke, and the number of cylindersare “loaded” into the on-screen ComponentSelection Box. These values are subse-quently used in the simulation process ofpredicting horsepower and torque. TheBore/Stroke menu choices should be con-sidered a “handy” list of common enginecylinder-bore and crankshaft-stroke values,not a description of engine configurations(e.g., V8, V6, straight 6, V4, etc.), materialcomposition (aluminum vs. cast iron), thetype of cylinderheads (hemi vs. wedge) orany other specific engine characteristics.The Bore/Stroke menu only loads bore ,stroke , and the number of cylinders intothe program database.

Bore, Stroke, & Compression Ratio

After making a selection from the Bore/Stroke menu, or when the individual bore,stroke, and number of cylinders have beenentered manually, the swept cylinder vol-ume (in cubic centimeters) and the totalengine displacement (in cubic inches) willbe calculated and displayed in the on-screen Component Selection Box. Theswept cylinder volume measures the vol-ume displaced by the movement of a singlepiston from TDC (top dead center) to BDC(bottom dead center). This “full-stroke”volume is one of the two essential elements

Choices loaded and calculatedafter Bore/Stroke Menu selection

When a selection ismade from the Bore/Stroke menu, the bore,stroke, and the numberof cylinders are “loaded”into the on-screenComponent SelectionBox. In addition, theswept cylinder volume(in cubic centimeters)and the total enginedisplacement (in cubicinches) will be calculatedand displayed.


required in calculating compression ratio.We’ll discuss compression ratio in moredetail later in this guide, but for now it’shelpful to know that compression ratio isdetermined with the following formula:

Compression Ratio =

Swept Cyl Vol + Combustion Space Vol———————————————————

Combustion Space Vol

In other words, the total volume that existsin the cylinder when the piston is locatedat BDC (this volume includes the SweptVolume of the piston and the CombustionSpace Volume) is divided by the volumethat exists when the piston is positioned atTop Dead Center. For the time being, it’simportant to keep in mind that your selec-tions of bore and stroke dimensions greatlyaffect compression ratio. When the stroke,and to a lessor degree the bore, is in-creased while maintaining a fixed combus-tion-space volume, the compression ratiowill rapidly increase. And, as is the case inMotion’s simulation software, if the com-pression ratio is held constant—because itis a menu selection and, therefore, fixedby the user—the combustion space vol-ume must increase to maintain the desiredcompression ratio. This may seem moreunderstandable when you consider that ifthe combustion-space volume did not in-crease, a larger swept cylinder volume (dueto the increase in engine displacement)would be compressed into the same finalcombustion space, resulting in an increasein compression ratio.

One example of how bore and strokecan have a significant affect on compres-sion ratio is demonstrated in a destrokedracing engine. Engine designers have re-alized for many years that shorter-strokeengines waste less power on pumping-work(more on this later), leaving more horse-power available for rotating the crankshaft.Furthermore, if engine displacement is heldconstant (by a concomitant increase in boresize), a larger cylinder diameter will ac-commodate larger valves, and increasingvalve size is one of the most effective waysto improve breathing and horsepower. Allof these changes taken together can pro-

duce considerable power increases, pro-viding the compression ratio is maintainedor increased. Unfortunately, it is difficult tomaintain a high compression ratio in shortstroke engines because: 1) Overall sweptvolume is often reduced, 2) combustionspace volume is proportionally larger, and3) short stroke engines move the pistonaway from the combustion chamber moreslowly requiring increased valve-pocketdepth to maintain adequate piston-to-valveclearance.

It is easy to use your PC to simulate a1-inch stroke, 5-inch bore, 8-cylinder en-gine of 157 cubic inches with an 11:1 com-pression ratio that produces over 500hp at8000rpm (with the power continuing toclimb rapidly!). Unfortunately, it may benearly impossible to build this engine withmuch more than 9 or 10:1 compressionbecause the volume needed for the com-bustion chamber, valve pockets, and headgasket on a 5-inch bore is large comparedto the swept volume produced by the short1-inch stroke. If we installed typical small-block race heads on this short-stroke con-figuration and great care was used inmachining the valve pockets, a total com-bustion space of about 55cc’s might bepossible. Unfortunately, this would onlyproduce 7:1 compression. The entire com-bustion-space volume would have to beless than 32 cubic centimeters to generatea compression ratio of 11:1 or higher!

Bore And Stroke Vs. Friction

You know that selecting stroke lengthwill determine, in part, the cubic-inch dis-placement of the engine. And from theprevious section you now realize that ob-taining high compression ratios with shorterstroke engines can be quite difficult. Whatyou may not realize is that setting the strokelength determines, to great extent, theamount of power lost to friction. The strokefixes the length of the crank arm, and thatdetermines how fast the piston and ringpackages “rub” against the cylinderwalls atany given engine speed. And here’s an-other rub (pardon the pun): 70% to 80% ofall IC engine frictional losses are due topiston and ring-package “drag” against the


cylinderwall! If a 1-inch stroke engine run-ning at 8000rpm looses 10 horsepower dueto cylinderwall friction, the same engine willloose 25hp with a 2-inch stroke, 40hp witha 3-inch stroke, 70hp with a 4-inch stroke,90hp with a 5-inch stroke, and 120hp witha 6-inch stroke at the same 8000rpm crankspeed. That means this 6-inch stroke en-gine consumes 110 more horsepower thanits 1-inch stroke counterpart just to drivethe pistons and rings up and down theirbores!

Try this simulation on your PC to dem-onstrate frictional losses. Build a 603.2cubic-inch engine, we’ll call it a very “Long-Arm” smallblock, with the following compo-nents:Bore: 4.000 inchesStroke: 6.000 inchesCylinderheads: Smallblock/Stock

Ports And ValvesValve Diameters: 2.02 Intake; 1.60

Exhaust (Valvediameters must bemanually selectedto disable the“Auto Calculate”function. Thiskeeps the valvesize fixed whenthe bore size ischanged—requiredfor the next test.)

Compression Ratio: 10.0:1Induction Flow: 780 CFMIntake Manifold: Dual Plane

Exhaust System: Small-TubeHeaders withOpen Exhaust

Camshaft: Stock Street/Economy

Lifters: HydraulicThis combination produces the power andtorque curves shown in the accompanyingchart (Long-Arm). Note that the powerdrops to zero above 5500rpm. Translated,this means that at 5500rpm the engine isusing all the power it produces to over-come internal friction!

Now build the same displacement en-gine, but change the bore and stroke com-bination to:Bore: 5.657 inchesStroke: 3.000 inchesThis “Short-Arm” configuration displaces thesame 603.2 cubic inches, but it producesover 100hp at 5500rpm and maintains a“non-zero” power level until about 6500rpm.Where was the “extra” horsepower hiding?The majority is “freed” by lower pistonspeeds and reduced bore-wall friction fromthe 3-inch shorter stroke.

Motion simulations use an empiricalequation to calculate frictional mean effec-tive pressure (Fmep):

Fmep = (12.964 + (Stroke x (RPM /6))) x (0.0030476 + (Stroke x (RPM /6) x 0.00000065476))

Once the Fmep is known, the horsepowerconsumed by friction can be calculated with

This “Short-Arm”configuration (solidlines) displaces thesame 603.2 cubic inchesas the “Long-Arm” test(dotted lines), but itproduces nearly 100hpmore at 5500rpm andmaintains a “nonzero”power level until about6500rpm. Where was the“extra” horsepowerhiding? The majority is“freed” by lower pistonspeeds and reducedbore-wall friction.


the following equation:

Friction HP =(Fmep x CID x RPM) / 792,000

Applying these equations to the two previ-ous test engines reveals the following:

Power consumed by friction with the 6-inch stroke engine @ 5000rpm: 120.7hpPower consumed by friction with the 3-inch stroke engine @ 5000rpm: 44.8hpSo far we’ve discussed how stroke

length affects displacement, compressionratio, and frictional losses within the en-gine. In addition to these effects, bottom-end configuration can have a measurableeffect on power consumed by the pro-cesses of drawing fresh fuel/air mixture intothe engine and forcing burned exhaustgasses from the cylinders (a process calledpumping, consuming what is termed pump-ing work). However, the IC engine is aremarkably complex mechanism that cansometimes fool nearly anyone into believ-ing that they have discovered subtle, “se-cret” power sources. The logic behind thediscovery may seem to make perfectsense, but eventually turn out wrong ormisleading. This is the case with the beliefpurported by some “experts” that: 1)Shorter-stroke/larger-bore engine configu-rations have the potential to be more effi-

cient high-speed air pumps, and 2) Shorterstrokes are the key to producing high-speedhorsepower. Let’s take an in-depth look athow bore, stroke, and rod length affectpiston speed and pumping work.

Fallacy One:Stroke Length Vs. Pumping Work

The concept of pumping work is moreeasily understood when you remember thatthe IC engine is, at its heart, an air pump.Air and fuel are drawn into the cylindersduring the intake cycle, and burned exhaustgasses are pumped from the cylindersduring the exhaust cycle. The power re-quired to perform these functions is calledpumping work and is another source of“lost” horsepower.

The intake stroke begins with the pistonaccelerating from a stop at TDC. As itbegins to move down the bore, the sweptvolume within the cylinder increases, cre-ating a lower pressure. This drop in pres-sure causes an inrush of higher pressureair and fuel from the intake system to com-pensate or “fill” the lower pressure in thecylinder. As the piston continues to accel-erate down the bore, the speed of the in-duced charge also increases, until at about70-degrees after top dead center, the pis-ton reaches maximum velocity. At this point,

TDC

STROKE

BDCBDC

Piston Speed And Frictional LossesIncrease As Stroke Increases

The stroke length deter-mines, to great extent, theamount of power lost tofriction. 70% to 80% of allIC engine frictional lossesare due to piston and ring-package “drag” against thecylinderwall! A longerstroke increases the lengthof the crank arm, and thatincreases the speed ofpiston and ring contactwith the cylinderwall.


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PISTONTRAVEL

PISTONTRAVEL

1/2STROKE

1/2STROKE

BOREDIAMETER

Equal Displacement EnginesSweep Out Equal Cylinder Volumes

During Each Crank Degree

Small BoreLong Stroke

Large BoreShort Stroke

the greatest pressure drop exists within thecylinder, drawing outside charge with thegreatest force. It is this “force” that is thekey element in understanding pumpingwork. The differences in pressures thatcause the inrush of air and fuel are not afreebie from nature; they consume work.The greater the difference in pressurebetween the intake manifold and the cylin-der, the more pumping work will be re-quired to move the piston from TDC toBDC. Any increase in pumping work di-rectly reduces the power available at thecrankshaft.

Let’s consider various techniques thatcan reduce pumping work on the intakestroke and potentially increase power out-put. First of all, pumping work can be sub-stantially reduced by lowering cylinder vol-ume. A smaller cylinder takes less work tofill it. Unfortunately, reducing cylinder vol-ume also reduces the amount of air andfuel that can be burned on the powerstroke, lowering power output more thanany possible gains from a reduction inpumping work. So if power is the goal, re-ducing pumping work must be accom-

While it’s true thatshorter-stroke engines

consume less frictionalhorsepower, it is nottrue (as believed bysome) that shorter-

stroke engines, withtheir slower piston

speeds, generate lesspumping work. For equal

displacements, equalcylinder volumes are

swept out at eachincrement of piston

travel from BDC to TDC.Short-stroke/large-bore

engines pump just asmuch air as long-stroke/

small-bore engines.

plished without decreasing the swept vol-ume of the cylinder.

The question is how can we fill the samespace using less work? This can be trans-lated to read: How can we move the sameor more air/fuel volume into the cylinderwhile inducing less pressure drop betweenthe cylinder and the intake system? Solu-tions to this problem are widely used byracers and include larger intake valves,freer-flowing ports and manifolds, largercarburetors or injector systems, and tuned-length intake runners. All these techniquesallow the same or more air/fuel mixture toflow into the cylinder while consuming lesspumping work.

While it’s true that shorter-stroke enginesconsume less frictional horsepower to drivethe pistons up and down the cylinders (asdescribed in the last section), it may alsoseem logical that a shorter stroke engine,with its slower piston speeds, can reducepumping work. Here’s how this concept isoften described and how the underlyingthread of truth often remains undiscovered:First, picture a 6-inch stroke, small-boreengine. When the piston begins to move


away from TDC, the long stroke sends itquickly towards a maximum velocity (maxi-mum volume change) that occurs about70 degrees ATDC. At this point, the longstroke has accelerated the piston to veryhigh speeds that generate a strong pres-sure drop in the cylinder. It is claimed thatthis low pressure generates flow velocitiesso high that the cylinderheads, valves, andthe induction system become a significantrestriction to flow. The rapid buildup in flowand high peak flow rates lower pumpingefficiency, and the overall picture getsworse as engine speed and piston speedincrease. Now consider a 1-inch stroke,large-bore engine. Because the stroke ismuch shorter, peak piston speed—andtherefore peak pressure drop—are believedto be considerably lower. The reasoningcontinues that induction flow never reachesas high a rate but is spread out more evenlyacross the entire intake cycle. The shorterstroke is believed to allow the inductionsystem to more efficiently fill the cylinder,consuming less pumping work. As enginespeed increases, this improved efficiencylets the engine produce higher power lev-els at greater speeds than a longer-strokeengine.

The accompanying Chart-A (drawn froma simulation based on the underlying phys-ics) illustrates the actual differences in in-cremental flow between an engine with a

long stroke and the same displacementengine with a “destroker” crank and a largerbore. Notice how the shorter stroke enginedevelops less peak flow and spreads thesame total flow volume slightly more evenlyacross the entire intake stroke. This reduc-tion in peak flow allows the induction sys-tem to operate more efficiently, slightlyimproving cylinder filling and reducingpumping work. So that would seem to proveit, right? The chart shows that a shorterstroke pumps more efficiently. While thatmay seem to be the case, take a look atChart-B . Here’s a plot of the same twoengines, but this time the rod lengths havebeen adjusted so that they have exactlythe same rod ratio, that is, the length of allconnecting rods are now equal to 1.7 timesthe length of the strokes. With identical rodratios, the cylinders in both engines pumpexactly the same volume at each degreeof crank rotation. The “proven benefits” ofa short stroke disappear.

The misunderstanding that longer strokeengines are less efficient breathers prob-ably occurs because during dyno tests ofvarious combinations, identical rod ratiosare not maintained when stroke lengths arechanged. So measured differences inpower are mistakenly attributed to changesin stroke, rather than the real cause: Varia-tions in rod ratio. If you have any doubtsabout this, take a look at the simulation

While pumping work isnot directly affected by

bore or stroke, largerengines, like this 600cid

big block, generatetremendous pumping

work at high enginespeeds. The pumping

losses are due to restric-tions at the valves, ports

and runners, andthroughout the intake and

exhaust systems. Atsome point in engine

speed, there no longerexists enough time to

move the gasses throughpassages of fixed sizes.

When this happens,power takes a nose-dive.


This graphillustrates thedifferences inflow as strokechanges. Noticehow the shorterstroke enginedevelops lesspeak flow andspreads the sametotal flow volumeslightly moreevenly across theentire intakestroke.

Here’s a plot ofthe same two

engines, but thistime the rod

lengths have beenadjusted to

produce the same1.7 rod ratio. With

identical rodratios, the cylin-

ders in bothengines pump

exactly the samevolume at eachdegree of crank

rotation.

The graph showsthe flow per crankdegree for 3-inchstroke vs. 24-inchstroke engines.Both engineshave the same603cid and sweepout the samevolume at each 2-degree incrementof crank rotationthroughout theentire intakestroke, despite the21-inch differencein stroke lengths !


depicted in Chart-C . Here’s a test of a 3-inch stroke, 5.67-inch bore engine com-pared to a very long 24-inch stroke enginehaving a 2-inch bore. Both engines havethe same total displacement of 603ci andthe same rod ratios (2.0 times the strokelength). The graph shows that the pistonsin both engines sweep out the same vol-ume at each 2-degree increment of crankrotation throughout the entire intake stroke,despite the 21-inch difference in strokelengths!

So it’s not stroke, but rather rod ratio,that has subtle, measurable effects onpumping work and peak flow rates. Thelonger the rod ratio, the more spread outthe induction flow and, potentially, thegreater the high-speed power. The shorterthe rod ratio, the higher the peak flow andthe earlier in the intake cycle the flow peakis reached. Considering these flowchanges, short-rod combinations couldshow potential gains on engines that ben-efit from a strong carburetor signal at lowerengine speeds, like short-track and streetengines. But even rod ratio cannot beconsidered in isolation. When the rod lengthis changed it also affects lateral loads onthe pistons, and it can change the way thepistons “rock” in their bores and the waythe rings “flutter” or seal against the cylin-der walls. So, depending on the lubricantsused, the cylinder block and piston con-figurations, and several additional factors,the “theory” of rod ratio and the powerchanges that are predicted may or maynot be found on the dyno. At least younow have an understanding of the under-lying theory, even though it makes its pre-dictions in isolation of many other interre-lated variables.

While we’re on the subject of rod ratio,many readers may be wondering why thisvariable was not included in one of thepull-down menus in Motion Software’sengine simulation. There are two reasonsfor this: 1) As just stated, the effects ofvarying rod length are very subtle and of-ten masked by other variables within theengine (such as piston side loads, bore-wall friction, etc.) making accurate model-ing extremely difficult, and 2) The subtlechanges in swept volume primarily require

wave-action dynamics for rigorous analy-sis (a process discussed in the completeDeskTop Dynos book available from Mo-tion Software). So rather than add a com-ponent that could not be modeled as accu-rately as the other variables in the pro-gram, rod length is not presented as a “tun-able” element (although it is included withinthe simulation process).

Fallacy Two: Long-Stroke TorqueVs. Short-Stroke Horsepower

Another remarkably widespread beliefhas to do with stroke length vs. enginespeed and torque vs. horsepower. A greatmany performance enthusiasts believe thatlong-stroke engines inherently developmore low-speed torque and short-strokeengines generate more high-speed horse-power. This understanding (probably as theresult of reading various magazine articlesover the years) is almost completely incor-rect, and some enthusiasts are quite defi-ant when confronted with the fact thatstroke, by itself, has little to do with high-speed or low-speed power potential. But,like the fallacy that short-stroke enginesare more efficient air pumps, there is athread of truth lying hidden under the sur-face.

To start off, long- and short-stroke en-gines of equal displacement will producethe same cylinder pressures during thepower stroke for the same swept volumes(assuming that they have consumed equalquantities of air and fuel; and we’ve al-ready demonstrated that stroke, by itself,has almost nothing to do with induction ef-ficiency). While the pistons in a longer-stroke engine are smaller in diameter and,therefore, experience less force from thesame cylinder pressure (force on the pis-ton is directly related to the surface area ofthe piston and pressure in the cylinder),the crank arms are longer and have agreater mechanical advantage. The result,believe it or not, is that equal displacementengines of unequal stroke lengths—expe-riencing the same cylinder pressures—willproduce the same torque at the crankshaft.Stroke, however, is not an isolated vari-able; it affects the design of many other


engine components. It’s in this interrelat-edness that we find the roots of misunder-standing about stroke vs. horsepower.

Probably the most direct reason for thebelief that longer-stroke engines are lower-speed “torque generators,” not capable ofproducing as many horsepower per cubicinch, is that many longer-stroke engineshave smaller bores. An engine with smallerbores almost always has smaller combus-tion chambers, and smaller combustionchambers have smaller valves. So mostlonger-stroke engines are forced, by thedesign of the cylinderheads and the entireinduction system, to produce less power athigher engine speeds. It’s not the length ofthe stroke that limits power potential, itsreduced flow from smaller valves, ports,and runners.

For more concrete proof, refer back tothe simulation we performed in the earliersection on stroke vs. friction. In that test,we compared identical displacement en-gines of 603ci, one with a 6-inch stroke(dotted lines on graph) to one with a 3-inch stroke (solid lines). Both engines usedthe same size valves and the same 780cfm

induction flow capacity. As previously dem-onstrated, the increase in horsepower fromthe shorter-stroke engine is due to a re-duction in bore-wall friction, adding about100 horsepower at 5500rpm. But look atthe shape of the solid-line curves. Theymatch the shape of the dotted-line curvesfor the longer-stroke engine. There is nonoticeable boost in low-speed torque fromthe engine with a 6-inch stroke. A steadily-increasing power loss from additional cyl-inderwall friction is clearly visible, but no-tice that the short-stroke engine producedno less torque between 2000 and 2500rpm,a range that many believe the longer-strokeengine should easily out-torque its short-stroke counterpart.

From a “real-world” mechanical stand-point, longer-stroke engines have draw-backs that limit their high rpm potential. Ascan be seen in our simulation, bore-wallfriction becomes a substantial power rob-ber. Ring seal also becomes a seriousproblem. As the stroke increases, higherand higher piston speeds cause the ringsto “flutter” against the cylinderwalls decreas-ing their sealing ability. It’s also possible

SMALLERVALVES

SMALLERPORTS

SMALLERRUNNERS

SMALLERBORE

TOP VIEW

Smaller-Bore/Longer-Stroke EnginesTypically Have Smaller Valves and Runners

And Produce Peak Power AtLower Engine Speeds

Many people believethat longer-stroke

engines are lower-speed“torque generators,” notcapable of producing as

many horsepower percubic inch as shorter-

stroke engines of equaldisplacement. What is

often overlooked is thatmost engines with

smaller bores almostalways have smaller

combustion chambers,smaller valves, and

smaller runners. It’s notthe length of the stroke

that limits powerpotential, its reduced

flow through the induc-tion system.


that at very high piston speeds during thepower stroke, the piston moves so quicklythat the rings can’t maintain a seal betweenthe bottom of the ring and the ring land inthe piston, increasing blowby and furtherdecreasing horsepower. The mechanicalloads on the piston and rod assembly inlong-stroke engines also become a seri-ous factor. As the piston is accelerated fromTDC down the bore, extremely high ten-sion loads are imparted to the rods andpistons, and added component weight tocompensate for the additional loads fur-ther increases stress.

While there are good reasons whylonger-stroke engines are ill suited to high-rpm applications, don’t confuse these an-cillary problems with the basic design rela-tionships between bore and stroke. Anddon’t fall into the trap of believing, like thou-sands of enthusiasts, that selecting a longerstroke will automatically boost low-speedpower.

THE CYLINDERHEADAND VALVE DIAMETER MENUS

The Cylinderhead pull-down menu, lo-cated just to the right of the Bore/Strokemenu, contains two sub-menus that allowthe simulation of various cylinderhead de-signs and a wide range of airflow charac-teristics. The first submenu, Head/PortDesign , lists general cylinderhead charac-teristics, including restrictive ports, typicalsmall- and big-block ports, and even 4-

valve cylinderheads. Each category in-cludes several stages of port/valve modifi-cations from stock to all-out race. A selec-tion from this menu is the first part of atwo-step process that Motion simulationsuse to accurately model cylinderhead flowcharacteristics. This initial selection deter-mines the airflow restriction generated bythe ports. That is, a selection from the firstsubmenu fixes how much less air than thetheoretical maximum peak flow will passthrough each port. What determines peakflow? That’s selected from the secondCylinderhead submenu: Valve Diameters .Valve size fixes the theoretical peak flow(called isentropic flow—more on that later).Most cylinderheads flow only about 50%to 70% of this value. The Valve Diametersubmenu allows the direct selection of valvesize or Auto Calculate Valve Size maybe chosen, directing the simulation to cal-culate the valve diameters based on boresize and the degree of cylinderhead port-ing/modifications.

While it may seem as if the variousCylinderhead menu choices simply refer toranges of airflow data stored within theprogram, that is not the case. If each menuselection fixed the flow capacity of thecylinderheads to a specific range of values(like typical flow bench data measured at astandardized pressure drop), the simula-tion would be severely limited. While asimulation based on this technique mightadequately calculate mass flow for enginesthat used nearly identical cylinderheads, ac-

The Head/Port Designmenu, under the Cylin-derhead main menu, listsgeneral cylinderheadcharacteristics, includingrestrictive ports, typicalsmall- and big-blockports, and even 4-valvecylinderheads. Eachcategory includesseveral stages of port/valve modifications fromstock to all-out race.


curacy would diminish rapidly for engineswith even modestly larger or smaller portsand valves.

There are several reasons why the de-termination of port flow in sophisticatedengine simulations can not be based strictlyon flow-bench data. First of all, flow gen-erated in the ports of a running engine isvastly different than the flow measured ona flow bench. Airflow on a flow bench is asteady-state, measured at a fixed pressuredrop (it’s also dry flow, but a discussion ofthat difference is beyond the scope of thisbook). A running engine will generate rap-idly and widely varying pressures in theports. These pressure differences directlyaffect—in fact, they directly cause—the flowof fuel, air, and exhaust gasses within theengine. An engine simulation program cal-culates these internal pressures at eachcrank degree of rotation throughout thefour-cycle process. To determine mass flowinto and out of the cylinders at any instant,you need to calculate the flow that occursas a result of these changing pressuredifferences. Since the variations in pres-sure, or pressure drops, within the engineare almost always different than the pres-sure drop used on a flow bench, flow benchdata cannot directly predict flow within theengine.

Before we delve deeper into the differ-ences between flow-bench data and themass flow calculated by an engine simula-tion program, let’s permanently put to restthe idea of storing the needed airflow data

The Valve Diametermenu allows the direct

selection of valve size oryou can select Auto

Calculate Valve Size ,directing the simulation

to calculate the valvediameters based on bore

size and the degree ofcylinderhead porting and

other modifications.

within the program. Consider for a momentan engine simulation program that’s basedon internally-stored airflow data. If youassume that the program is capable ofmodeling a wide a range of engines (likeMotion Software simulations), then it mustalso store a wide range of cylinderheadflow data. Flow would have to be recordedfrom a large number of engines, startingwith single-cylinder motors and working upto 1000+ cubic-inch powerplants. If it werepossible to accumulate this much data,there are additional shortcomings in alookup model. Since the pressures insidethe IC engine are constantly changing, alookup program would have to containranges of flow data measured from zero tothe maximum pressures differentials devel-oped within the running engine. Even if thismuch data could be assembled and filedaway, additional data for each engine andcylinderhead would be needed to predictthe flow changes when larger valves wereinstalled; and even more data would beneeded to model port modifications. It soonbecomes clear that the sheer bulk of flowdata needed for a comprehensive lookupmodel would probably consume morespace than exits on your hard drive (andthat’s considering that most of us now owngigabyte and larger drives)!

Cylinderhead ChoicesAnd Discharge Coefficients

While it is impractical to base an engine


simulation on extensive flow-bench data,measured cylinderhead flow figures are,nonetheless, commonly used in sophisti-cated engine simulations. Rather thanstored in extensive “lookup” tables withinthe programs, flow-bench data can be usedas a means to compare the measured flowof a particular port/valve configurationagainst the calculated isentropic (theoreti-cal maximum) flow. The resulting “ratio,”called the discharge coefficient , hasproven to be an effective link between flow-bench data and the simulated mass flowmoving into and out of the cylindersthroughout a wide range of valve openingsand pressure drops. Furthermore, the dis-charge coefficient can be used to predictthe changes in flow for larger or smallervalves and for various levels of port modi-fications. In other words, it’s the dischargecoefficient, not flow bench data, that pro-vides a practical method of simulating massflow within a large range of engines. Sincethis is such a powerful and often misun-derstood concept, the following overviewshould prove helpful in understandingwhat’s happening “behind the scenes”

when various choices are made from theCylinderhead pull-down menu.

The choices in the Cylinderhead menuare purposely generic. This can be frus-trating if you are interested in modeling onlyone engine or a single engine family. Inthese cases it would be ideal to list theexact components you wished to test, inname or part-number order. But if you areinterested in simulating different engines,including popular 4-, 6-, and 8-cylinderpowerplants, the choices provided in theCylinderhead menu (and many of the othermenus) offers considerable modelingpower. The menu choices move from re-strictive heads to smallblock, big block, andfinally to 4-valve-head configurations. Ifeach of these choices loaded a higher ab-solute flow curve into the simulation, theywould cover only a very narrow range ofengines. Instead, each of the menu choicesmodel an increasing flow capacity and re-duced restriction. This makes it possible toaccurately simulate a lawnmower engine,a high-performance motorcycle engine, anall-out Pro Stock big block, or a mild streetdriver, each having different port and/or

PRESSURE

CFM

SIMULATEDPRESSURE-WAVE FLOW

FLOW BENCHSTEADY-STATE FLOW

Airflow on a flow bench is steady state, but Motion's engine simulation program cal-culates internal pressures and mass flow at each degree of crank rotation throughoutthe four-cycle process. Since the variations in pressure, or pressure drops, within theengine are rarely equal to the pressure drop used on a flow bench, flow bench datacannot directly predict flow within the engine.


valve sizes, using selections from the samemenu!

The basis for this “universality” is thateach menu selection uses a different dis-charge-coefficient curve rather than airflowcurve. While the discharge coefficient datais derived from flow-bench data, it is di-mensionless (has no length, weight, mass,or time units) and is applicable to any cyl-inderhead of similar flow efficiency. Toclarify this concept, picture two large roomsconnected by a hole in the adjoining wall.When pressure is reduced in one room, airwill flow through the hole at a specific rate.It is possible to calculate what the flowwould be if there were no losses from heat,turbulence, etc. This flow rate, called theisentropic flow, is never found in the realworld, but is, nevertheless, a very usefulterm. It’s the rate of flow that “nature” willnever exceed for the given pressure dropand hole size. If we measure the actualflow through the hole and divide it by thecalculated isentropic flow, we will havedetermined the flow efficiency or dischargecoefficient of the hole:

Discharge Coefficient (always less than 1)= Measured Flow / Calculated Flow

Since every orifice has some associatedlosses, the discharge coefficient is alwaysless than 1. If the calculated discharge

coefficient for our hole in the wall was0.450, this would mean that the hole flows45% as much air as an ideal hole. In otherwords, it’s 45% efficient.

Let’s see how this concept can be ap-plied to cylinderheads. When an intakevalve of a specific size is opened a fixedamount, it exposes a flow path for air calledthe curtain area. Through this open area,measured in square inches just like thehole in the wall, air/fuel mixture moves ata specific rate depending the pressure dropacross the valve (as you recall, the pres-sures are calculated by the simulation soft-ware at each degree of crank rotation). Ifthe valve and port were capable of perfectisentropic flow, the simulation equationscould calculate the precise mass flow thatentered the cylinder during this moment intime. But real cylinderheads and valves arefar from perfect, and it’s flow bench datathat “tells” the simulation how far fromperfect the real parts perform. By dividingthe isentropic flow by the flow-bench data(both at the same pressure drop) of acomparable head/port configuration at eachincrement of valve lift, the simulation soft-ware creates a discharge coefficient curvethat it can use as a correction factor forport flow at all other pressure drop levels.

The true power of this method lies inthe fact that the cylinderhead tested on theflow bench—used to develop the discharge-

Theoretical Isentropic Flow

Actual Measured Flow 45% Of Isentropic Flow

The absolute maximumflow rate through an

orifice with no lossesfrom turbulence, heat, or

friction is called theisentropic flow. This is

the flow rate that“nature” will never

exceed for the givenpressure drop and holesize. If we measure theactual flow through the

hole and divide it by thecalculated isentropic

flow, we will havedetermined the flow

efficiency or dischargecoefficient .


coefficient curve—may have been designedfor an entirely different engine and useddifferent valve sizes. Nevertheless, as longas the head configuration matches thesimulated cylinderhead, in other words, theports have similar flow efficiency, the dis-charge-coefficient curve will closely adjustisentropic flow to real-world corrected flow.So a specific choice in the Cylinderheadmenu, say “Pocket Porting With LargeValves” may accurately model a factoryperformance head for a smallblock Chevy,a cylinderhead on a 4-cylinder Toyota en-gine, or even a motorcycle head. With dis-charge coefficient corrections, it’s not theabsolute flow numbers that are important,it’s how well the valve and port flow fortheir size that really counts.

As you made selections from the Cylin-derhead menu, you may have wonderedwhether the flow data used to “correct” theisentropic flow in any of the menu itemswould match the published flow of cylin-derheads that you would like to simulate.Considering what you now know about therelationships between cylinderheads ofsimilar efficiencies and how flow data isused by the program, it becomes clear thatknowing the flow-bench data used in theprogram would probably not be helpful. Theflow may not have been obtained from thesame valve sizes or even the same brand

VALVEHEAD

VALVESEAT

CURTAIN AREA

When a valve of a specific size is openeda fixed amount, it exposes a flow pathcalled the curtain area . Through this openarea (measured in square inches) gassesmove at a specific rate depending on thepressure drop across the valve.

cylinderheads. On the other hand, it isentirely feasible for the simulation to sub-stitute user-entered flow data from testsconducted on a specific set of cylinder-heads. This data could then be used todevelop custom discharge-coefficientcurves that would, in turn, closely modelcylinderheads. While this is not currentlysupported, it will be incorporated in upcom-ing versions of Motion simulation software(if you haven't already, send in your regis-tration card; you’ll receive information onupgrades and new releases).

Ram Tuning AndPressure Waves

Up to this point our discussion has cen-tered around airflow, valve size, and dis-charge coefficients. The assumption hasbeen that reducing restriction and increas-ing flow efficiency will allow more air/fuelmixture to enter the cylinders and producemore horsepower. Initially this is true, butwhen port and valve sizes are increased,power begins to fall off at lower speeds,then at higher speeds, and finally very largepassages reduce power throughout theentire rpm range! At first, this may seemquite mysterious. If minimizing restrictionwas the key to improved airflow and power,this phenomenon would indeed be inexpli-cable. But as we’ve discovered, the ICengine does not function by simply direct-ing the flow of air and fuel as a hose pipedirects the flow of water. Powerful wavedynamics take control of how air, fuel, andexhaust gasses move within the inlet andexhaust passages. Because of these phe-nomena, there is an optimum size for theports and valves for any specific applica-tion, from street economy to all-out dragracing.

Consider what happens after the intakevalve opens and the piston begins moving


down the bore on the intake stroke. Aspressure drops, air/fuel mixture enters thecylinder. During this portion of the induc-tion cycle, any restriction to inlet flow re-duces cylinder filling and power output. So,while the piston moves from TDC to BDC,the rule of thumb for the ports and valvesis “the bigger the better.” After the pistonreaches BDC and begins to travel back upthe bore on the early part of the compres-sion stroke, the intake valve remains openand the cylinder continues to fill with air/fuel mixture. This “supercharging” effect,caused by the momentum built up in themoving column of air and fuel in the portsand inlet runners, adds considerable chargeto the cylinder and boosts engine output.However, as soon as the piston begins tomove up the bore from BDC, it tries topush charge back out of the cylinder. Largerports and valves not only offer low restric-tion to incoming charge, but they make iteasier for the piston to reverse the flowand push charge back out of the cylinder.Furthermore, a low restriction inductionsystem has a large cross-sectional areaand allows the incoming charge to movemore slowly (in feet per second), so thecharge carries less momentum and, onceagain, is more easily forced back out ofthe cylinder. So the rule of thumb to opti-

When the piston movesfrom TDC to BDC on theintake stroke, “the biggerthe better” is the rule forports and valves. Afterthe piston reaches BDCand begins to travel backup the bore, it tries topush charge back out ofthe cylinder. Now the ruleof thumb for power is“smaller ports and highairflow speeds.” Acompromise must befound. This race enginefinds that critical balancewith large runners thatgenerate 700-ft/sec peakport velocity at very highengine speeds.

mize power for the period of time betweenBDC and intake-valve closing is “smallerports that produce higher airflow speedsare better.” Since it’s not possible to rap-idly change the size of the ports as theengine is running, a compromise must befound that minimizes restriction and opti-mizes charge momentum.

As it turns out, optimum port and valvesizes for performance applications at aspecific engine speed must be smallenough to allow the charge to reach speedsof about 700 feet per second during theintake stroke. Then, as the piston movesfrom BDC to the intake valve closing point,the momentum generated by these speedscontinues to force air and fuel into thecylinder, optimizing charge density. Typi-cally, the best port size and cam timing forperformance allows a slight “reversion” offresh charge just before the intake valvecloses.

Unfortunately, the smaller port cross-sectional areas required to generate opti-mum flow velocities create a restriction toairflow and increase pumping work. If portcross-sectional areas were smaller and flowvelocities increased much beyond 700 feet/second, the added restriction and increasedpumping work would reduce overall cylin-der filling and engine output would suffer.


The “magic” balance found at about 700feet per second between charge momen-tum, inlet restriction, and pumping workallows the cylinders to fill with the greatestmass of air/fuel mixture throughout thecomplete induction cycle from IVO (intakevalve opening) to IVC (intake valve clos-ing).

What is the best cross-sectional areafor any specific engine; the area that gen-erates about 700 feet/second peak flowvelocities? That is a very tough question toanswer. In fact, optimum port shapes andcross-sectional areas are so interrelatedwith other engine variables, like enginedisplacement and rpm, cam timing, exhaustsystem configuration, intake manifold de-sign, compression ratio, and more, thatengine simulation software is needed tosort through the complexity and find the“magic” combinations. Furthermore, a fullwave-action modeling program, like Dyno-

mation discussed in the complete Desk-Top Dynos book, is required for this tasksince it’s the dynamics of finite-amplitudewave motion that are responsible for therecorded changes in horsepower.

Despite these complexities, there aresome “rules of thumb” that can be helpfulin selecting a workable port for common

Minimum Port Cross-sectionalArea Equal 0.85 Times ValveArea For Street Applications

Minimum Port Cross-sectionalArea Equal 0.90 Times ValveArea For Race Applications

Despite the interrelated-ness of engine speed,

cam timing, exhaustsystem configuration,

and more, there are some“rules of thumb” in

selecting a workable portfor common applications.

Cylinderheads willprobably provide good

performance if the areaof minimum port cross-

section is equal to 0.9times the valve area for

race applications and0.85 for street use.

700 Feet Per SecondMaximum Flow

Speed

The best cross-sectional area for anyspecific engine, one that generates about

700 feet/second peak flow velocity, isdifficult to find. Optimum port shapes and

cross-sectional areas are so interrelatedwith engine displacement and rpm, cam

timing, exhaust system configuration,intake manifold design, compression ratio,and more, that engine simulation software

is needed to sort through the complexityand find the “magic” combinations.


applications. If the area of minimum crosssection in the ports is equal to 0.9 timesthe areas of the valves for race applicationand 0.85 times the valve areas for streetuse, the heads will probably provide goodperformance. This rule applies to typicalautomotive engines, not to high-rpm mo-torcycles or low-rpm aircraft. However, fortypical powerbands between 5000 and8500rpm, this rule should be reasonablyclose, providing the valve diameters arenot too small or too large for the applica-tion.

These same relationships between portcross-sectional areas and valve diametersare used in Motion Software’s Filling AndEmptying simulation. As you move downthe Cylinderhead menu, choices withineach section—say within the smallblock orbig block categories—the minimum portcross-section increases from 0.85 to 0.9times the selected valve areas. However,since the current simulation does not in-corporate as robust a wave-action analy-sis as the Dynomation program (discussedin the complete DeskTop Dynos book), portcross-sectional areas are not tunable bythe user.

The phenomenon of cylinder filling bycharge momentum is often called ram tun-ing , and is commonly thought to be an

independent property of induction tuning,separate from the complex theory of wavedynamics that most professional racers onlyvaguely recognize. Because the conceptof ram tuning is easy to grasp, it has beenwidely “applied” for many years. Long in-jector stacks, extremely large ports andvalves, and other measures have beenused to facilitate the free flow of additionalcharge into the cylinder after BDC on theintake stroke. But now that we have un-covered the fine balance needed betweenport cross-sectional area, charge momen-tum, airflow velocity, pumping work, camtiming, and engine speed, it is more obvi-ous that large ports and valves, by them-selves, are not the answer. Unfortunately,the insatiable desire to have large portsand valves seems to drive the cylinder-head market. Most customers simply wantlarger ports and valves. So head manufac-turers and porters turn out droves of headswith ports that are too large for the appli-cation. While they reduce restriction duringthe intake stroke, low charge momentumduring the majority of the rpm range of theengine reduces overall cylinder filling andpower output.

One important thing to learn from thisbook should be to permanently discard thenotion that “bigger is better” when it comes

The phenomenon ofcylinder filling by chargemomentum is oftencalled ram tuning, andoptimizing this effectmeans finding a balancebetween port cross-sectional area, airflowvelocity, pumping work,cam timing, and enginespeed. This custommanifold was designed—with the help of simula-tions, including Dynoma-tion (discussed in thecomplete DeskTopDynos book) to find thatillusive balance onsmallblock Fords.


to ports and valves. The right combinationis the key.

Sorting Out CylinderheadMenu Choices

Now that we have covered some of thebasic theory behind the choices in theCylinderhead menus, here’s some practi-cal advice that may help you determinethe appropriate selections for your applica-tion.

Low Performance Cylinderheads—There are three “Low Performance” cylin-derhead selections listed at the top of theHead/Port Design submenu. Each of thesechoices is intended to model cylinderheadsthat have unusually small ports and valvesrelative to engine displacement. Heads ofthis type were often designed for low-speed, economy applications, with littleconcern for high-speed performance. Early260 and 289 smallblock Ford and to a les-sor degree the early smallblock Chevycastings fall into this category. Thesechoices use the lowest discharge coeffi-cient of all the head configurations listed inthe menu. Minimum port cross-sectionalareas are 85% of the valve areas or some-what smaller and, if Auto Calculate Valve

Size has been selected (more on this fea-ture in the next section), relatively small(compared to the bore diameter) intake andexhaust valve diameters will be used.

The first low-performance choice mod-els an unmodified production casting. Thesecond choice “Low Performance/PocketPorting” adds minor porting work performedbelow the valve seat and in the “bowl” areaunder the valve head. The port runnersare not modified. The final choice “LowPerformance/Ported, Large Valves” incor-porates the same modifications and in-cludes slightly larger intake and exhaustvalves. Valve size increases vary, but theyare always scaled to a size that will gen-erally install in production castings withoutextensive modifications.

The low-performance choices havesome ability to model flathead (L-head &H-head) and hybrid (F-head) engines.While the ports in these engines are con-siderably more restrictive than early Fordsmallblock engines, by choosing Low-Per-formance and manually entering the exactvalve sizes, the simulation will, at least,give you an approximate power output thatyou can use to evaluate changes in camtiming, induction flow, and other variables.Also, it’s not essential that your model pro-duce an absolutely accurate horsepower

The “Low Performance”cylinderhead choices are

intended to model cylinder-heads that have unusually

small ports and valves.Heads of this type wereoften designed for low-

speed, economy applica-tions, with little concern for

high-speed performance.Early 260 and 289 small-

block Ford and to a lessordegree the early smallblockChevy castings fall into this

category.


number. A great deal of useful informationcan be found by simply looking at thechanges in power that result from variouscombinations of parts.

Smallblock Cylinderheads— The small-block and big-block choices comprise thetwo main cylinderhead categories in theHead/Port Design submenu. Choices fromthese two groups apply to over 90% of allperformance engine applications from mildstreet use to all-out competition.

The basic smallblock selections modelheads that have ports and valves sizedwith performance in mind. Ports are notexcessively restrictive for high-speed op-eration, and overall port and valve-pocketdesign offers a good compromise betweenlow restriction and high flow velocity. Thestock and pocket-ported choices are suit-able for high-performance street to modestracing applications.

The next step is “SmallBlock/FullyPorted, Large Valves” and this cylinderheadmoves away from street applications. Thiscasting has improved discharge coeffi-cients, greater port cross-sectional areas,and increased valve sizes. Consider thishead to be an extensively modified, high-performance, factory-type casting. It doesnot incorporate “exotic” modifications, likeraised and/or welded ports that requirecustom-fabricated manifolds. “SmallBlock/

Fully Ported, Large Valves” heads are high-performance castings that have additionalmodifications to provide optimum flow forracing applications.

The last choice in the smallblock groupis “SmallBlock/Race Porting And Mods.”This selection is designed to model state-of-the-art, high-dollar, Pro-Stock type cyl-inderheads. These custom pieces are de-signed for one thing: Maximum power. Theyrequire hand-fabricated intake manifolds,have excellent valve discharge coefficients,and the ports have the largest cross-sec-tional areas in the smallblock group. Thishead develops sufficient airflow speeds forgood cylinder filling only at high engine rpm.

Big Block Cylinderheads— All big-block selections are modeled around headswith “canted” valves. That is, the valvestems are tilted toward the ports to im-prove the discharge coefficient and overallairflow. All ports have generous cross-sec-tional areas for excellent high-speed per-formance.

The first three choices are based on anoval-port design. These smaller cross-sec-tional area ports provide a good compro-mise between low restriction and high flowvelocity for larger displacement engines.The stock and pocket-ported selections aresuitable for high-performance street tomodest racing applications.

The “SmallblockCylinderhead” menuchoices model cylinder-heads that have portsand valves sized withperformance in mind,like the heads on thisLT1. The stock selec-tions are not exces-sively restrictive forhigh-speed operation,and overall port andvalve-pocket designoffers a good compro-mise between lowrestriction and high flowvelocity.


The final two selections simulate exten-sively modified rectangular-port heads.These choices are principally all-out, big-block Chevy heads, however, they closelymodel other extremely aggressive high-performance racing designs, like theChrysler Hemi head. As with the smallblockcategory, the “Big Block/Fully Ported, LargeValves” heads are not suitable for moststreet applications. These castings havehigh discharge coefficients, large portcross-sectional areas, and increased valvesizes. This head is basically a factory-typecasting but extensively improved. However,it does not incorporate “exotic” modifica-tions, like raised and/or welded ports thatrequire custom-fabricated manifolds.

The last choice in the big-block group is“Big Block/Race Porting And Mods.” Thisselection is designed to model state-of-the-art, high-dollar, Pro-Stock cylinderheads.These custom pieces, like their smallblockcounterparts, are designed for maximumpower. They require hand-fabricated intakemanifolds, have optimum valve dischargecoefficients, and the ports have the largestcross-sectional areas in the entire Head/Port Design submenu, except for 4-valveheads (discussed next). These speciallyfabricated cylinderheads only develop suf-ficient airflow for good cylinder filling withlarge displacement engines at very highengine speeds.

4-Valve Cylinderheads— The last threeselections in the Head/Port Designsubmenu model 4-valve cylinderheads.These are very interesting choices sincethey simulate the effects of very low-re-striction ports and valves in stock andperformance applications. The individualports in 4-valve heads begin as single, largeopenings, then neck down to two Siamesedports, each having a small (relatively) valveat the combustion chamber interface. Sincethere are two intake and two exhaust valvesper cylinder, valve curtain area is consid-erably larger than with the largest single-valve-per-port designs. In fact, 4-valveheads can offer more than 1.5 times thecurtain area of the largest 2-valve heads.This large area, combined with high-flow,low-restriction ports greatly improves airand fuel flow into the cylinders at highengine speeds. Unfortunately, the portsoffer an equally low restriction to reverseflow (reversion) that occurs at low enginespeeds when the piston moves up thecylinder from BDC to Intake Valve Closing(IVC) on the final portion of the intakestroke. For this reason, 4-valve heads, evenwhen fitted with more conservative portsand valves, can be a poor choice for small-displacement, low-speed engines. On theother hand, the outstanding flow charac-teristics of the 4-valve head put it in an-

The “Big-BlockCylinderhead” selectionsare modeled aroundheads with canted valves.Ports have generouscross-sectional areas.The first three menuchoices model oval-portdesigns. The final twoselections simulatemodified rectangular-portheads. The appropriateselection for this L29 big-block Chevy wouldprobably be the secondor third menu choice—thefourth menu choicemodels a head with flowcapacity beyond thestock L29.


other “league” when it comes to high horse-power potential on large engines at highengine speeds.

The first choice in the 4-valve group is“4-Valve Head/Stock Ports And Valves.”This simulates a 4-valve cylinderhead thatwould be “standard equipment” on a fac-tory high-performance or sports-car engine.These “mild” heads offer power comparableto high-performance 2-valve castingsequipped with large valves and pocketporting. However, because they still haverelatively small ports, reasonably high portvelocities, and good low-lift flow character-istics, they often show a boost in low-speedpower over comparable 2-valve heads.

The next choice, “4-Valve Head/PortedWith Large Valves” represents a mild re-work of “stock type” 4-valve heads. Largervalves have been installed and both theintake and exhaust flow has been improvedby pocket porting. However, care has beentaken not to increase the minimum cross-sectional area of the ports. These changesprovide a significant increase in power withonly slightly slower port velocities. Rever-sion has increased, but overall, these headsshould show a power increase throughoutthe rpm range on medium to large displace-ment engines.

The final choice, “4-Valve Head/Race

Porting And Mods,” like the other “RacePorting And Mod” choices in the Head/PortDesign submenu, models an all-out racingcylinderhead. This selection has the great-est power potential of all. The ports areconsiderably larger than the other choices,the valves are larger, and the dischargecoefficients are the highest possible. Theseheads suffer from the greatest reversioneffects, especially at lower engine speedson small displacement engines. (Theseheads, like all head choices, are “scaled”to engine size, so that smaller engines au-tomatically use appropriately smallervalves—providing the Auto Calculate ValveSize option is selected—and smaller ports.)If you would like to know what “hidden”power is possible using any particular en-gine combination, try this cylinderheadchoice. It is safe to say that the only wayto find more power, with everything elsebeing equal, would be to add a super-charger, nitrous-oxide injection, or useexotic fuels.

When To Choose SmallblockOr Big-Block Cylinderheads

Making appropriate Head/Port Designchoices may be easier now that you areaware of some of the less-obvious issues.

The “4-ValveCylinderhead” menu

selections modelcylinderheads with 4-

valves per cylinder.These heads can offer

more than 1.5 times thecurtain area of the

largest 2-valve heads.This large valve area,

combined with high-flow,low-restriction ports

greatly improves air andfuel flow into the

cylinders at high enginespeeds. These Cosworthheads were designed for

the English Ford V6.When they were raced in

England several yearsago, they regularly beat

Chevy V8s.


However, there are additional aspects ofmenu selections that should be emphasizedto avoid confusion when modeling specificcylinderheads. Perhaps the most “confus-ing” issue surrounds the selection of headsfrom the smallblock group for engines thatare commonly recognized as big blocks.

Many big-block engines use cylinder-heads that are simply “scaled-up” small-block designs. The valve and port sizes onthese heads may be 10 to 20% larger thantheir smallblock counterparts, but that’sabout the only difference. The ports remaina tall, rectangular shape, valve stem anglesrelative to the port centerline are identicalor nearly identical, and combustion cham-ber shapes are often the same. When pro-portionally larger heads are installed onengines of proportionally larger displace-ment, overall flow efficiency remains verysimilar. So for engines like the big-blockChrysler wedge, Olds, Pontiac and severalFord big blocks, plus many other engines,a selection from the smallblock group,

This is another exampleof a smallblock Chevy

with canted valves andspread-port intakes.

Developed for the limitedIMSA GTP racing pro-gram in the early ’90s,

these Chevrolet splayed-valve heads should be

correctly modeled by thethird or fourth big-block

menu choice.

This smallblock Chevyhead from Brodix incorpo-rates canted valves and aspread-port intake design,meeting the criterion for a“big-block” for simulationpurposes. The second orthird menu choices (oval-port) are probably the bestmodels for these heads.

combined with entering (or having the pro-gram calculate) larger valve diameters, willclosely simulate the power levels of these“big-block” engines.

Some big blocks have cylinderheads ofsubstantially different design. These “true”big block heads are built for higher perfor-mance levels and have visibly differentports, valves, and/or combustion chambers.Big-block Chevy heads fall into this cat-egory with two common configurations. Thefirst is a milder, “street,” oval-port versionand the second uses larger rectangular-ports. The larger head is, without question,a high-performance, high-rpm design, evenon large-displacement engines. Both ofthese heads “cant” the valve stems towardthe centerline of the port. This shifted valveposition improves the discharge coefficientand allows for slightly larger valves.

This same design philosophy is used inChrysler’s Hemi heads. An all-out racingdesign, the Hemi “cants” the valves evenfurther and uses combustion chambers


shaped in a true hemisphere. The portshave large cross-sectional areas with highdischarge coefficients. These heads, likeall big-block heads with the most “aggres-sive” designs, come up short on low-speedperformance because of poor cylinder fill-ing and high reversion, but produce ex-ceptional horsepower at high enginespeeds on large displacement engines.

Finally, there are engines that havesmallblock displacements but use cylinder-heads that look like they came straight offof a big block. Many 4- and 6-cylinder“sports-car” engines fall into this category,using single- or double-overhead cams,canted valves, and even “hemi” combus-tion chambers. More rarely, smallblockChevys even turn up with big-block heads.Chevrolet’s maximum performance“splayed-valve” heads for the smallblockChevy IMSA racing program were releasedin the early nineties. These cylinderheadsare simply a “mini” big-block design, andmodeling of this head should be done usingthe big-block choices in the Head/PortDesign submenu.

Valve Diameters AndThe “Auto Calculate” Feature

The Valve Diameters submenu is lo-cated in the lower half of the CylinderHead

The last two choices inthe big-block menu willalso model Chrysler's426-Hemi cylinderhead.The last menu choice willsimulate all-out ProStockconfigurations.

menu. The first selection is “Auto Calcu-late Valve Size.” This powerful feature in-structs the simulation software to determinethe most likely valve sizes to be used withthe current engine based on an assess-ment of the bore diameter and the Head/Port Design selection. The Auto-Calculatefunction is active when selected and bydefault when the simulation program isstarted or whenever “Clear Current Com-ponent Selections” is chosen from the Util-ity menu. Auto Calculate is especially help-ful if you are experimenting with a severaldifferent bore and stroke combinations orcomparing different engine configurations.Auto Calculate will always select valves ofappropriate size for the cylinderheads un-der test and it will never use valve sizesthat are too large for the current bore di-ameter. If you are using version 2.5 orgreater of Motion Software’s simulation, theprogram will display the calculated valvediameters in the Component Selection Boxalong with “(Auto),” indicating that the valvesizes were automatically determined. Ear-lier versions simply displayed “(Auto)”;calculated sizes were not shown.

While the Auto-Calculate feature is help-ful during fast back-to-back testing, it maynot “guess” the precise valve sizes used,and therefore, not simulate power levelsas accurately as possible. In these situa-


tions refer to the second part of the ValveDiameters menu. Here you will find a listof exact valve sizes consisting of commonintake and exhaust combinations, or youcan choose “Other...” from the bottom ofthe menu and directly enter any valve di-ameters within the acceptable range of theprogram (approximately 0.800- to 2.75-inches). When exact valve sizes are se-lected by either of these methods, the di-ameters are displayed in the ComponentSelection Box along with “(Man),” indicat-ing that the sizes were manually entered.When “(Man)” is active, the program dis-ables auto-calculation of valve diameters

and will not automatically change the dis-played values, regardless of the cylinder-heads or cylinder bore diameters chosenfor the test engine. You can change valvediameters at any time by simply choosingdifferent sizes from the menu or by select-ing “Other...” again. You can also re-en-able the Auto Calculate feature at any timeby re-selecting it from the Valve Diametersubmenu.

THE COMPRESSION RATIO MENU

C/Ratio is the third pull-down menu lo-cated in the main menu bar. A selection

Auto Calculate willalways select valves ofappropriate size for thecylinderheads undertest. While this feature ishelpful during fast back-to-back testing, it maynot “guess” the precisevalve sizes. In thesesituations refer the list ofexact valve sizes, or youcan choose “Other...”from the bottom of themenu and directly entervalve diameters.

��

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COMBUSTIONCHAMBERVOLUME

COMBUSTIONSPACE

VOLUME

PISTONATTDC

While combustion-chamber volumeis simply the volume in the

cylinderhead, the combustion-spacevolume is the total enclosed volume

when the piston is located at TDC.This space includes the volume inthe combustion chamber, plus any

volume added by the piston notrising to the top of the bore, lessany volume due to the piston or

piston dome protruding above thetop of the bore.


from this menu establishes the compres-sion ratio for the simulated engine from8:1 to 16:1 (version 2.5.7—under develop-ment when this book was published—sup-ports a compression ratio range of 6:1 to18:1 using the “Other” selection). The com-pression ratio is a comparison of the geo-metric volume that exists in the cylinderwhen the piston is located at BDC (bottomdead center) to the “compressed” volumewhen the piston reaches TDC (top deadcenter). As you recall, compression ratio iscalculated with the following formula:

Compression Ratio =

Swept Cyl Vol + Combustion Space Vol———————————————————

Combustion Space Vol

Let’s take a close look at this relation-ship to discover exactly what compressionratio is and how compression ratio worksinside the IC engine.

Compression Ratio Basics

The above compression-ratio equationcontains two variables: 1) swept cylindervolume, and 2) combustion space volume.These volumes are the only two variablesthat affect compression ratio. However,each of these variables is made up ofmultiple elements, so the first step in ex-ploring compression ratio must be to “dis-sect” these variables.

Swept cylinder volume is the most

The C/Ratio establishesthe compression ratio

for the simulated engine.The compression ratio is

a comparison of thegeometric volume that

exists in the cylinderwhen the piston is

located at BDC (bottomdead center) to the

“compressed” volumewhen the piston reaches

TDC (top dead center).Changing compressionratio has a pronouncedeffect on engine power.

straightforward of the two. As you discov-ered in the previous Bore/Stroke sectionof this guide, swept volume is calculatedby Motion’s Filling And Emptying simula-tion—and displayed in the Component Se-lection Box—as soon as the bore andstroke have been selected for the testengine. Swept volume is simply the three-dimensional space displaced by the pistonas it “sweeps” from BDC to TDC, and isdetermined solely by the bore diameter andstroke length.

The other variable in the compression-ratio equation is combustion-space volume.This is the total volume that exists in thecylinder when the piston is located at TDC.This space includes the volume in thecombustion chamber, plus any volumeadded by the piston not rising fully to thetop of the bore, less any volume due to thepiston protruding above the top of the bore.The complexity involved in combustion-space volume can be a stumbling block forsome enthusiasts. You may find it helpfulto refer to the accompanying drawings forillustrations of these concepts.

A good way to visualize these variablesis to imagine yourself a “little man” wan-dering around inside the engine. Let’s takea walk inside the combustion space. Pic-ture what it would look like in the cylinderwith the piston at TDC. You would see thecombustion chamber above you like aceiling. Your floor would be the top of thepiston. If the piston (at TDC) didn’t risecompletely to the top of the cylinder, around


the edges of the floor you would see a bitof the cylinderwall, with the head gasketsandwiched between the head and block,like a low molding around the room. Theremay be notches in the top of the pistonjust under your feet (don’t trip!). If the pis-ton had a dome, it might look like a smallretaining wall rising from the floor, to per-haps knee high. The combustion spacewould be larger if the piston was positionedlower down the bore or if the notches un-der your feet were deeper, and it would besmaller if the retaining wall (dome) volumewas larger. This entire space is “home” forthe compressed charge when the pistonreaches TDC. This is the volume thatmakes up the denominator of the compres-sion-ratio calculation. Now let’s continue our“ride” in the cylinder, but this time picturewhat it looks like when the piston is posi-tioned at BDC. The very same volumesthat we just described (chamber, dome,notches, gasket, etc.) are still there, butare now located well above our head. Itlooks like the room has been stretched,like the elevator ride in the Haunted Houseat Disneyland. This “stretched” volume iswhat is described in the numerator of theequation. It’s simply the original combus-tion volume plus the volume added by the“sweep” of the piston as it traveled fromTDC to BDC. The ratio between these twovolumes is the compression ratio.

Changing Compression Ratio

PISTONAT

TDC

PISTONAT

BDC Picture yourself as a“little man” inside thecylinder. With the pistonat BDC, the very samevolumes that exist atTDC are still there, butnow the swept-volume ofthe cylinder has beenadded. The ratio betweenthese two volumes is thecompression ratio.

A quick look at the compression-ratioequation reveals that if engine displace-ment (swept volume) is increased, eitherby increasing the bore or stroke, the com-pression ratio will rise. In fact, with every-thing else being equal, a longer stroke willincrease compression ratio much moreeffectively than increasing bore diameter.This is due to the fact that a longer strokenot only increases displacement, but ittends to decrease combustion space vol-ume since the piston moves higher the bore(in our “little man” example, raising the floorcloser to the ceiling). This “double posi-tive” effect results in rapid increases in com-pression ratio for small increases in strokelength. On the other hand, increasing cyl-inder bore diameter also increases com-pression ratio but much less effectively.This is caused by the increase in combus-tion volume that accompanies a larger bore(again, using our “little man,” a larger boreadds more floor space by increasing thediameter of the room), partially offsettingthe compression-ratio increase that occursfrom increasing cylinder displacement.

Changing combustion space, the otherelement in the equation, will also alter thecompression ratio. Anything that reducesthe combustion volume, while maintainingor increasing the swept volume of the cyl-inder, will increase the compression ratio.Some of the more common methods aredecreasing the volume of the combustion


chambers (by replacing or milling theheads), using thinner head gaskets, chang-ing the location of the piston-pin or rodlength to move the piston closer to thecombustion chamber, installing pistons withlarger domes, and others. All of thesemodifications will increase compressionratio. And that increases horsepower, right?Well, surprisingly, the answer to that ques-tion is often “Yes!”. But do you know why?

Why Higher Compression RatiosProduce More Horsepower

Most automotive enthusiasts believe thata higher compression ratio directly raiseshorsepower output. Before you read thenext few paragraphs, see if you can ex-plain to yourself why compressing thecharge before combustion is necessary. Oris it necessary? If it is, what causes a boostin power? Does it burn the fuel more com-pletely or efficiently? What about the powerthat’s needed to compress the charge? Isit simply “lost” energy?

The answers to these questions lie inthe laws of thermodynamics. Luckily, wedon’t have too delve to deeply into thiscomplex subject to gain an insight into howcompression before ignition works. Let’stake a simplified look at what happensinside the engine when the compressionratio is increased. Picture a spherical com-bustion space containing twice as muchvolume as the cylinder that’s attached to it.This configuration produces an engine witha very low 1.5:1 compression ratio. Justbefore the intake valve closes, the pistonis positioned at BDC, and the cylinder andthe spherical volume are exposed to atmo-spheric pressure of about 14psi. As thepiston moves up the bore, the valve closesand the charge is compressed. When thepiston reaches TDC the pressure in thecylinder will rise to about 21psi. At this pointthe spark plug fires and drives the post-combustion pressure to about 250psi andthe piston is pushed back down the bore.About 12 to 14 degrees after TDC, thecylinder pressure driving the piston downthe bore will be about 230psi.

Now picture the same engine, exceptthis time the spherical combustion space

has been reduced in size to about 1/10thof the volume of the cylinder. The com-pression ratio is now 10:1. As the intakevalve closes at BDC, the cylinder and thevolume are once again at atmosphericpressure. When the piston reaches TDCon the compression stroke, the small vol-ume has driven the compression pressureto about 400psi, or 18 times as high (thenew volume is about 18 times smaller).When the piston reaches 12 to 14 degreesafter TDC, the pressure on the piston isabout 1500psi. The higher compression ra-tio produces much higher cylinder pres-sures throughout the first half of the piston’stravel from TDC and BDC on the power

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COMPRESSIONRATIO1.5:1

COMPRESSIONRATIO10:1

A spherical combustion space containingtwice as much volume as the cylinderproduces a 1.5:1 compression ratio. Peakcylinder pressures (see text) will be about250psi. With a combustion space about1/10th of the volume of the cylinder, thecompression ratio is now 10:1. Peak pres-sures now reach about 1500psi. Thehigher compression ratio produced muchhigher cylinder pressures throughout thefirst half of the piston’s travel from TDCand BDC on the power stroke. This addi-tional pressure generates a much largerforce across the surface of the piston,and that increases torque and horse-power.


stroke. This additional pressure generatesa much larger force across the surface ofthe piston and that increases torque andhorsepower.

While it may now be clear that higherpost-ignition pressures result in increasesin torque output, you may still be wonder-ing how much power is consumed to com-press the charge to higher initial pressures.This is more easily understood if you pic-ture what happens when the spark plugdoesn’t fire. It certainly takes power to drivethe piston up the bore to compress thecharge, but without ignition nearly identicalpressures drive the piston back down thebore on the “power” stroke. The net result(forgetting about friction, mechanical, andheat losses) is zero. In other words, underideal circumstances it take no more netpower to compress a charge to a lazy 21psiwith a 1.5:1 compression ratio than it doesto raise it to 400psi with 10:1 compressionbecause the consumed work is returnedon the power stroke. In a real engine,higher compression ratios increase lossesfrom charge heating and, especially, fromincreased ring-to-cylinderwall friction. Butthe power lost is usually smaller than thepower gained. Testing done by GM manyyears ago indicated that, for gasoline asthe fuel, power will continue to increaseuntil compression ratios reach about 17:1.Considering that many racing engines arenow hovering around this level, GM mayhave been right.

Other Effects Of IncreasingCompression Ratio

As mentioned above, the greatest lossesfrom high compression and high cylinderpressures (except when detonation is in-duced) occur from ring/cylinderwall frictionand heat losses into the pistons and waterjackets. Beyond these well-known effects,changing the compression ratio has othersubtle consequences, some of which arelargely unexplored.

Let's examine a potential source of “hid-den” power. The combustion space volumeabove the piston tends to act as an ab-sorber, slightly smoothing the pressurepulses in the cylinder. This effect can

dampen the peaks of the negative pres-sure waves created when the piston movesdown the bore on the intake stroke. Whenthe compression ratio is increased, thecombustion space decreases. With lessvolume to absorb pressure pulses, the pres-sure drop on the intake stroke may be more“directly linked” to the induction system,improving cylinder filling. You can visual-ize this effect by picturing a small cylinderattached to a large room (the room exag-gerates the effects of the large combustionspace used in low compression ratio appli-cations). As the piston moves down thebore, any drop in pressure is dissipated inthe room, so that the carburetor—attachedat the adjacent wall—receives an extremelyweak signal and virtually no air/fuel mix-ture flows into the room or cylinder. Nowreduce the size of the room to a smallspace, more like the real-world conditionsin a high-compression engine. As the pis-ton moves down the bore, the drop in pres-sure is almost directly “linked” to the in-duction system, instantly drawing air andfuel into the cylinder. These same positiveeffects, though not as pronounced, shouldoccur anytime the compression ratio isincreased by decreasing the combustionspace.

Compression Ratio Assumptions

Changes in compression ratio havehundreds of consequences throughout theengine. However, Motion’s Filling AndEmptying simulation program analyzes onlysome of these changes. Changes thatoccur because of alterations in the wavedynamics within the engine are not wellpredicted by the current program. Thermo-dynamic effects, however, are very accu-rately modeled in the simulation. In orderto understand the variations in power andtorque curves produced as the compres-sion ratio is changed, it is helpful to under-stand some program assumptions.

The program assumes that ignition tim-ing is always optimum. That is, based onthe gasoline burn model used by the pro-gram, ignition occurs at a point that pro-duces peak cylinder pressures between 12to 14 degrees after top dead center. Tests


Here’s a source of“hidden” power. A large

combustion space thataccompanies a lowcompression ratio

(illustrated here by thelarge box attached to the

carburetor) tends to actas an absorber, smooth-

ing pulses in the cylin-der. This dampens the

negative peaks createdwhen the piston moves

down the bore on theintake stroke. When the

compression ratio isincreased, the combus-

tion space decreases.With less volume to

absorb pulses, the intakestroke is more “directlylinked” to the induction

system, improvingcylinder filling.

LARGE COMBUSTION CHAMBERWITH LOW COMPRESSION

SMALL COMBUSTION CHAMBERWITH HIGH COMPRESSION

have indicated that these same conditionsreproduced in the real world often optimizepower output.

The limitations of the combustion modelprevent changing the ignition point. Thesimulation of varying ignition timing—or per-forming sophisticated combustion analysisto reveal detonation, preignition, or provideemissions analysis—requires advancedtechniques. While these models do exist,they not only need full three-dimensionalmaps of the constantly-varying combustionspace, but they consume, literally, days ofcomputational time. Obviously, this is notpractical approach for a quick-response“what-if” program.

THE INDUCTION MENU

The fourth main component menu es-tablishes an Induction system for the simu-lated test engine. An induction system, asdefined in Motion’s simulation, is everythingupstream of the intake ports, including theintake manifold, common plenums (if used),carburetor/throttle main body, venturis (ifused), and opening to the atmosphere. TheInduction menu is divided into twosubmenus: an Airflow menu to select the

maximum-rated airflow that can passthrough the induction system, and theManifold submenu to establish an intake-manifold configuration. One choice fromeach of these two groups fixes a specificinduction system from among thousandsof possible combinations.

Airflow Selection

The first Induction submenu is used toselect the rated airflow for the inductionsystem. This menu consists of two 2-bar-rel carburetor selections, twelve 4-barrelcarburetor/fuel injection choices, and an“Other...” selection in which you can di-rectly specify the rated airflow from 100 to3000cfm. The first two selections “install”either a 300- or 500-cfm 2-bbl carburetoron the test engine. These are the only 2-barrel choices directly available in themenu. The remaining choices range from300 to 1100cfm on 4- or 8-barrel carbure-tor and fuel injection applications.

In order to perform “apple-to-apple” com-parisons among the Airflow selections, it isimportant to realize that the ratings for 2-barrel carburetors are measured at a pres-sure drop twice as high as the pressure


used to rate 4-barrel carburetors and fuel-injection systems. Rated airflow for 2-bar-rels is typically measured at a pressuredrop of 3 inches of mercury (the pressuredifferential maintained across the carbure-tor during airflow measurement at wide-open throttle). This is often written as “3-in/Hg” (“Hg” is the symbol for mercury usedin the periodic table of elements). Thehigher pressure drop increases the mea-surement resolution for smaller carburetorsand “shifts” the flow numbers toward therange commonly found in automotive ap-plications (roughly, 100 to 700cfm).

Knowing the differences in rating meth-ods, it is a simple task to convert any 2-bblflow into it’s 4-bbl equivalent. Here’s theformula:

2-bbl Flow4-bbl Flow = ——————

1.414

Using this conversion, it is possible to simu-late virtually any 2-barrel carburetor induc-tion system. For example, a custom 2-bar-rel that flows 650cfm at 3-in/Hg, would flow460cfm if measured at 1.5-in/Hg (you canconfirm this using the above formula). Bymanually entering 460cfm into the Compo-nent Selection Box of the simulation, theprogram will accurately model this custom2-barrel.

The remaining choices in the InductionAirflow menu are labeled with “4/8-BBLCarb Or Fuel Inj.” This means that the

selections below the first two designateairflow ratings that were measured at 1.5-in/Hg. The “4/8-BBL” indicates that the in-duction system can consist of single ormultiple carburetors that, combined, pro-duce the rated airflow. For example, themenu choice “1000 CFM 4/8-BBL Carb OrFuel Inj” can indicate one 1000cfm 4-bblcarburetor or two 500cfm 4-bbls. The same1000cfm selection could even indicate three470cfm 2-bbls (i.e., 3 x 470 = 1410cfm;converting to 4-bbl flow: 1410/1.414 =997cfm). The important thing to rememberabout airflow selection is that the programmakes no assumption about the type ofrestriction that makes up the carburetor orinjection system. The airflow is simply arating of the total restriction of the induc-tion system.

Motion’s Filling-And-Emptying modeluses this restriction value to calculate acritically important variable needed by thesimulation to accurately determine massflow into the cylinders: manifold vacuum.Here’s how the process works:1) The simulation runs through an entire

cycle (all four Otto cycles) to initially de-termine the total mass flow into thecylinders. This determination—becauseit is based on a degree-by-degree,crank-angle analysis—takes the entirerange of engine variables into consider-ation, including displacement, enginespeed, valve size, cam timing, compres-sion, and assumptions about the intakemanifold, exhaust system, and more.

The first Inductionsubmenu is used toselect the rated Airflowfor the induction system.This menu consists of 2-barrel and 4/8-barrelcarburetor/fuel injectionchoices, and an“Other...” selection bywhich you can directlyspecify the rated airflowfrom 100 to 3000cfm.


2) Since there can never be mass accu-mulation or loss within the engine, thesame mass that flows into the cylindersmust also flow through the induction(and exhaust) system.

3) Since atmospheric pressure exists onone side of the carburetor or throttlebody, using the selected induction re-striction (chosen from the airflowsubmenu), the program can calculate thevacuum generated in the manifold thatwill produce the same predicted rate.

4) When the manifold vacuum has beendetermined, the charge density can becalculated. Knowing the charge densityis the final step used by the program to“iterate,” or home-in on, a precision de-termination of the air/fuel mass enteringthe cylinder.Throughout these simulation steps, the

carburetor creates a restriction that pro-duces a vacuum in the manifold. Largercarburetors produce less restriction anddecrease vacuum. As vacuum drops andmanifold pressure approaches atmosphericlevels, the density of the charge becomesgreater. When the density of the air/fuelcharge increases, a greater mass of airand fuel can be drawn into the cylinders,resulting in higher power output. Seemingly,the conclusion here is that the greater theflow capacity of the carburetor, the morepower the engine can produce. While thismay hearken back to the old cliché “thebigger the better,” in theory, this trend isabsolutely correct. The carburetor can be

thought of as a density-regulating device,and any restriction generated by the car-buretor (or injection system) increasesmanifold vacuum and decreases chargedensity and power.

In the real world, however, carburetorsmust generate a pressure drop across theventuris in order to atomize fuel and air inthe proper proportions. In fact, most car-buretors must generate at least 0.5-in/Hgpressure drop at the lowest airflow levelsto function properly (e.g., during part throttletransition to full throttle at low enginespeeds). This same requirement does notexist for modern fuel-injection systems.Many high-performance electronic andmechanical injection systems maintain pre-cise air/fuel mixtures throughout the rpmrange while generating negligible restric-tion and manifold vacuum at full throttle.While there are practical limitations to “thebigger is better” rule, in theory, low restric-tion and high charge density clearly pro-duce more power.

Airflow Menu Assumptions

As higher airflow levels are selected fromthe Induction menu, the simulation lowersthe restriction within the induction system.This decrease in restriction increasescharge density. Along with this concept,the simulation assumes that the air/fuelratio is always at the precise proportion foroptimum power. While optimum air/fuel ra-tios are more achievable with fuel injection

Choosing “Other” fromthe Induction Airflowmenu allows direct entryof any airflow rating forthe simulated inductionsystem, providing it fallswithin the acceptablerange limits. Airflowvalues entered using thistechnique are assumedto be rated at a pressuredrop of 1.5-in/Hg, thestandard of measure-ment for 4-barrel carbu-retors.

“Other” direct airflow entryfrom Induction menu.

Notice range limits below.


systems, a carefully tuned carburetor alsocan come remarkably close to ideal fuelmetering. Regardless of whether the simu-lated engine uses carburetors or fuel injec-tion, the power levels predicted by theprogram can be considered optimum,achievable when the engine is in “peak”tune and the induction system is workingproperly.

The airflow (in Cubic Feet per Minute,or CFM) selected from the Induction Air-flow menu is also assumed to be the totalrated airflow into the engine. On dual-inletor multiple-carburetor systems, the totalairflow is the sum of all rated airflow de-vices. So a manifold equipped with twin1100cfm Holley Dominators would have arated airflow of 2200cfm. If an air cleaneris used, total airflow must be adjusted tocompensate for the increase in restriction(contact the element manufacturer or flowtest the carburetor/air-cleaner as an assem-bly).

Also keep in mind the unique way air-flow capacities are handled on IndividualRunner (I.R.) manifolds. On these induc-tion systems, each cylinder is connectedto a single “barrel” or injector stack with noconnecting passages that allow the cylin-ders to “share” barrels. The total rated flowfor these induction systems is dividedamong the number of cylinders. For ex-ample, a smallblock V8 equipped with 4

Weber carburetors (having 8 barrels) mayhave a total rated flow of 2000cfm. To prop-erly model this system, 2000cfm is directlyentered into the simulation by choosing“Other...” from the Airflow menu. When an“I.R.” manifold is selected from the secondpart of the Induction menu (more on mani-folds next) the airflow is divided into all 8cylinders, allotting 250cfm to each cylin-der.

Induction Manifold Basics

The lower half of the Induction menuconsists of five manifold choices. Each ofthese manifolds applies a unique tuningmodel to the induction system, but beforewe cover the particulars of each selection,it is helpful to review induction tuning andhow wave-dynamic models are used toreveal manifold function. For a more in-depth look at induction tuning and wavedynamics, refer to the complete DeskTopDynos book available from Motion Soft-ware, Inc.

The flow of air and fuel within enginepassages is influenced by waves gener-ated by rapidly changing pressures withinthe induction and exhaust systems. Thesepressure “pulses” arise from the release ofhigh pressure exhaust gasses into the portsand headers during the exhaust cycle andby the “pumping action” of the piston dur-

Larger carburetorsproduce less restrictionand decrease manifoldvacuum. Seemingly, thegreater the flow capacity,the more power theengine will produce. Intheory this trend iscorrect. The carburetoris a density-regulatingdevice with any restric-tion increasing manifoldvacuum and decreasingcharge density, resultinglower power. In the realworld, however, carbure-tors must generate apressure drop of about0.5-in/Hg to functionproperly.


ing the intake cycle. These pressure wavesare thousands of times stronger than waveswe are familiar with in everyday life: com-mon acoustic or sound waves. At thesehigh energy levels, IC engine pressurewaves, now referred to as finite-amplitudewaves, alter their shape as they travelthrough passages, and they interact witheach other in ways considerably morecomplex than simple acoustic waves. Thecombination of these phenomena make the“solutions” to finite-amplitude wave analy-sis extremely difficult.

This complexity is multiplied by the con-figuration of the ports, valves, manifoldrunners and plenum shapes. The levels ofcomplexity continue to grow as the intakeand exhaust valves open and close. Ini-

tially, with the intake valve closed, the port/runner system forces pressure wavesthrough a basic oscillation cycle, bouncingback and forth off of the closed valve atone end of the runner and the open mani-fold plenum at the other. This “reasonablysimple” process, called 1/4-wave or organ-pipe tuning, delivers a series of decaying-amplitude pressure pulses to the closedintake valves. When the intake valves areopen, however, the system switches fromorgan-pipe tuning to a much more com-plex Helmholtz resonator. This is the sameresonance that you can duplicate by blow-ing into the neck of a jug, creating a deep“whirring” sound. Not only does pressure-wave analysis have to deal with this morecomplex resonance, but the Helmholtz

When the intake valve isclosed, the port/runner

system forces pressurewaves through a basicoscillation cycle called1/4-wave or organ-pipe

tuning . When the intakevalve is open, the

system switches to amuch more complex

Helmholtz resonator withchanging volume as the

piston moves in thecylinder and varying

restriction as the valveopens and closes.

The lower half of theInduction menu consistsof five Manifold choices.Each manifold applies aunique tuning model to

the induction system.Refer to the text for

information on thedesign of each manifold,

an overview of how themanifold boosts poweror torque, and finally, a

description of theassumptions and

recommendationsassociated with each

menu selection.

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CLOSED VALVE:SIMPLE, 1/4-WAVE

ORGAN PIPE TUNING

OPEN VALVE:COMPLEX HELMHOLTZRESONANCE


resonator is changing its volume as thepiston moves in the cylinder, and the re-striction at the “neck” is also varying, asthe valve moves through its lift curve.Combining all these factors with the alreadycomplex interaction of finite-amplitudewaves gives a glimpse of the mathemati-cal sophistication needed to analyze pres-sure waves inside the IC engine.

After all this, you may be wondering howmuch of an influence this morass of com-plex pressure waves has on engine perfor-mance. The answer is a lot! They can ei-ther aid or restrict cylinder filling depend-ing on the design of inlet ducting. A care-fully constructed Pro-Stock induction sys-tem will use these invisible pressure wavesto gain hundreds of horsepower. Evenstreet engines can use induction tuning toimprove throttle response, fuel economy,and add “seat-of-the-pants” power.

Induction-tuning power benefits comefrom several techniques, but the most di-rect is to harness the suction wave cre-ated during the intake stroke. Optimum run-ner lengths will return a reflection of thisnegative pressure wave to the intake valveon the next intake cycle when the pistonreaches maximum velocity, about 70 de-grees after TDC. The returning suctionwave combines with the low pressure cre-ated by the piston rapidly moving down

the bore to produce a powerful draw of airand fuel into the cylinder. Then, again, justbefore the intake valve closes as the pis-ton is beginning to move up the cylinder,the induction system returns a positivepressure pulse to minimize or prevent “re-version” of charge back out of cylinder.

These tuning effects can be adjustedfor specific applications by changing run-ner length, volume, passage interconnec-tions, and plenum configuration. In otherwords, installing intake manifolds of differ-ent design can have dramatic effects onhow the pressure waves are used to assistcylinder filling and control engine power.The following section details each of themanifold choices provided in the Inductionmenu and explains the assumptions andlimitations associated with each design.

Manifold Selection AdviceAnd Design Assumptions

The complex interaction of pressurewaves within the induction system requirea rigorous mathematical analysis, involv-ing the Method Of Characteristics dis-cussed in the complete DeskTop Dynosbook. This advanced technique uses con-siderable computational time and does notlend itself to simulations designed for arapid “what-if” interaction with the user. Fur-

The five manifoldsincluded in Motion’s

simulation are, by nomeans, a comprehen-

sive list of all theintake manifolds

available for ICengines. However, the

five discrete designswithin the menu canbe applied to a widerange of manifolds.

This SLP TPI manifoldcan be simulated by

applying a largeairflow rating to the

Tuned-Port Injectionmodel.


thermore, extensive analysis of intake mani-fold configurations would require the inputof many detailed variables, such as runnerlengths, volumes, taper angles, plenumconfiguration and dimensions, and muchmore. These technical inputs would furthershift the emphasis away from ease of usetoward a dedicated research tool. However,in order to evaluate manifold differences,any program—including Motion’s Filling-And-Emptying simulation—must incorpo-rate some pressure-wave analysis to makeaccurate estimations of power and torquedifferences. Motion’s program uses a “mini-wave action” model that offers three ad-vantages for most users: 1) rapid calcula-tion times allowing fast back-to-back test-ing, and 2) the mini-model does not re-quire dimensional data entry for the portsand runners, making component selectionextremely simple, and 3) overall accuracyis quite good (within 10%) and trends inpower differences between the five mani-fold types are reliable. For those individu-als involved in engine development pro-grams or intake manifold research, the mini-model will not provide sufficient data reso-lution to make subtle design changes. How-ever, for nearly everyone else, the mini-model should provide a good compromisebetween speed, ease of use, and predic-tive accuracy.

The five manifolds included in Motion’ssimulation are, by no means, a compre-hensive list of all the intake manifolds avail-

able for IC engines. The list of five should,instead, be interpreted as five discretedesigns that approach the limits of resolu-tion of the mini-model within the program.If you are interested in a manifold with adesign that falls in between two menuselections, you can often use the “trend”method to estimate power for a hybriddesign. For example, run a test simulationusing manifold Type A, then study the dif-ferences in power attributed to manifoldType B. The changes will indicate trendsthat should give you insight into how ahybrid manifold Type A/B might perform.Because a rigorous analysis of pressurewaves is not performed by the current pro-gram, keep in mind that the data you ob-tain might not match real-world dyno datawith some combinations. In general, how-ever, the trends and overall accuracyshould be within 10%.

For each manifold described below, youwill find information about its basic design,an overview of how the manifold boostspower or torque, and finally, a descriptionof the assumptions and recommendationsassociated with that individual design.

Dual-Plane Manifold —Remarkably, thewell-known and apparently straightforwarddesign of the dual-plane manifold is, argu-ably, the most complex manifold on thelist. An intake manifold is considered tohave a dual-plane configuration when 1)the intake runners can be divided into two

The Edelbrock PerformerQ-Jet represents the

current state-of-the-art indual-plane manifold

designs. This manifold issaid to have a 2 nd degreeof freedom. This power-ful resonance multiplies

the force of the pres-sures waves, simulating

the effects of longrunners, boosting low-and mid-range power.


groups, so that 2) each group alternatelyreceives induction pulses and 3) the pulsesare spaced at even intervals. If all of thesecriteria are met, the manifold is said to havea 2nd degree of freedom, allowing it toreach a unique resonance causing theentire manifold and all the runners oscil-late in unison. During this period, pressurereadings taken throughout the manifold willbe in “sync” with one another. This power-ful resonance multiplies the force of thepressures waves, simulating the effects oflong runners. Since longer runners typicallytune at lower engine speeds, not surpris-ingly, the dual-plane manifold is mostknown for its ability to boost low-end power.

The divided plenum is another commonfeature of dual-plane manifolds that furtherboosts low-end power. Since each side ofthe plenum is connected to only one-halfof the cylinders (4-cylinders in a V8), eachcylinder in the engine is “exposed” to onlyhalf of the carburetor. This maximizes wavestrength and improves low-speed fuelmetering (these effects are much less pro-nounced with throttle-body fuel-injectionsystems). However, the divided plenum canbecome a significant restriction at higherengine speeds and limit peak horsepower.

The main benefits of the dual-plane areits low-speed torque boosting capability,compact design, and wide availability foruse with both carburetors and injectionsystems. However, not all engines are

capable of utilizing a dual plane. Typically,engines that do not have an even firingorder or have too many cylinders to gen-erate a resonance effect will not benefitfrom a dual-plane manifold. While there aresome exceptions, engines having 2 or 4cylinders work best with this manifold. Sincemost V8 engines are basically two 4-cylin-der engines on a common crankshaft,even-firing V8s also benefit from the reso-nance effects of the dual-plane manifold.Motion’s simulation does not prevent choos-ing a dual-plane manifold on engines thatwill not develop a full resonance effect. Forexample, you can install a dual-plane mani-fold on a 5-cylinder engine, but the re-sults—a low-end power boost—are notreproducible in the real world, since aneffective dual-plane manifold cannot be builtfor this engine. The simulation is best uti-lized by modeling dual-plane manifoldscombinations that already exist rather thantesting theoretical fabrications.

Many dual-plane manifolds are hybridsincorporating facets of other manifold de-signs. Especially common is the use of anundivided or open plenum typically associ-ated with single-plane manifolds. Thesehodgepodge designs are attempts at har-nessing the best features while eliminatingthe worst drawbacks of various designs.Sometimes, the combinations are success-ful, adding more performance without muchof a sacrifice in low-speed driveability. With

The basic differencebetween single- and

dual-plane manifolds areclearly illustrated here.

The dual-plane (left)divides the plenum inhalf, with the runners

grouped alternately byfiring order. Each

cylinder “sees” only one-half of the carburetor, transferring a strong signal to theventuris. This manifold design is said to have a 2 nd degree of freedom, allowing it toreach a unique resonance that makes its short runners boost low-speed power. The

single-plane manifold (right) has short, nearly equal-length runners with an openplenum, much like a tunnel ram but “laid” flat across the top of the engine. The

manifold has excellent high-speed performance, but a loss of 2 nd degree of freedomprevents full-manifold resonance. That reduces low-speed torque and degrades

driveability and fuel economy.


these designs, you can successfully usethe “trend” method described earlier toestimate engine torque and power. How-ever, there is no shortage of manifolds thatcan reduce power without giving anythingback in driveability or fuel economy. In fact,some of the worst designs are remarkablybad. It is impossible to determine which ofthese combo designs is better than othersusing the Filling And Emptying simulation.Only a simulation that perform models in-take passages, including the complex ef-fects of multicylinder interference, can per-form this analysis. Unless you can performactual dyno testing on these manifolds tofind what works and what doesn’t, it maybe worthwhile to stick with more “plain-vanilla” designs that produce predictable

results.

Single-Plane Manifold —In a very realsense, a single-plane manifold as used onmost V8 engines is simply a low-profiletunnel ram. The tunnel-ram manifold (dis-cussed next) is a short-runner system com-bined with a large common plenum; adesign that optimizes power on all-out rac-ing engines where hood clearance is notan issue. The single-plane manifold com-bines short, nearly equal-length runnerswith an open plenum, but “lays” the entireconfiguration flat across the top of theengine. The results are quite predictable.A loss of 2nd degree of freedom preventsfull-manifold resonance. That produces aloss of low-speed torque, and depending

Many dual-plane mani-folds are hybrids. ThisEdelbrock dual-planemanifold is designed forthe 440 Chyrsler engineand has a partially openplenum. In this case, theopening adds mid-rangeand high-speed perfor-mance with little sacrificein low-speed driveability.Not all hybrid designsare as successful as thisone. In situations whereyou are not familiar withengine or manifoldcharacteristics, it may beworthwhile to stick with“plane-vanilla” designs.

A single-plane manifold issimply a low-profile tunnelram. The design combinesshort, nearly equal-length

runners with an openplenum, but “lays” the

entire configuration flatacross the top of the

engine. The single-planemanifold combines

improved flow capacity,higher charge density,

and short runners to buildsubstantial horsepower at

higher engine speeds.


on the size of the plenum and runners,single-plane manifolds can also degradedriveability and fuel economy. Furthermore,a large-volume, undivided plenum contrib-utes to low-speed problems by presentingevery cylinder to all barrels of the carbure-tor, lowering venturi signal and low-speedfuel metering accuracy (again, this draw-back is minimized with a fuel-injection sys-tem). On the other hand, the single-planemanifold (like the tunnel ram) combines im-proved flow capacity, potentially highercharge density, and short runner lengthsto build substantially more horsepower athigher engine speeds.

As a high-performance, high-speedmanifold, the single-plane design has manyadvantages, but it’s compact, low-profiledesign has drawbacks too. The runners areconnected to a common plenum like spokesto the hub of a wheel. This arrangementtends to create unpredictable interferenceeffects as pressure pulses moving throughthe runners meet in the plenum and stir upa complex soup. Large plenum volumeshelp cancel some these effects, but open-plenum, single-plane manifolds may pro-duce unexpected changes in fuel distribu-tion and pressure-wave tuning with spe-cific camshafts, headers, or cylinderheads(to some degree, these effects are presentin all manifold designs). Predicting thesewill-o’-the-wisp anomalies requires rigorousmodeling, well beyond the capabilities ofthe Filling-And-Emptying simulation. Cur-rently, pinning down these problems re-quires dyno testing with exhaust tempera-

ture probes to measure fuel distributionaccuracy.

Designers and engine testers have ex-perimented with hybrid single-plane mani-fold designs that incorporate various dual-plane features. One common modificationis to divide the plenum into a pseudo dual-plane configuration. While this does in-crease signal strength at the carburetor,uneven firing does not allow 2nd degree offreedom resonance. This modification cancause sporadic resonances to occurthroughout the rpm range with unpredict-able results. Spacers between the carbu-retor and plenum are also commonly usedwith single-plane manifolds often with posi-tive results, particularly in racing applica-tions. Spacers probably increase power fortwo reasons: 1) By increasing plenum vol-ume they tend to reduce unwanted pres-sure-wave interactions, and 2) A largerplenum improves airflow by reducing theangle the air/fuel must negotiate as it tran-sitions from “down” flow from the carbure-tor to “side” flow into the ports. While thereis no way to use trend testing to evaluatethe effects of a divided plenum, spacerscan be partially simulated. The increase inplenum volume tends to mutate the single-plane manifold into a “mini” tunnel ram, sohorsepower gains tend to mimic thoseobtained by switching to a tunnel ram de-sign (i.e., small performance improvements,when found, usually occur at high rpm).

Since the single-plane manifold typicallyreduces low-speed torque and improveshigh-speed horsepower, it is often the best

The typically compact,low-profile design of thesingle-plane manifoldhas drawbacks. Therunners are connected toa common plenum. Thisarrangement tends tocreate unpredictableinterference effects aspressure pulses movingthrough the runnersmeet in the plenum andstir up a complex soup,sometimes creatingirregular fuel-distribu-tion.


compact manifold design for applicationswhere engine speed is typically 4000rpmor higher. If the engine commonly runsthrough lower speeds, a dual-plane, indi-vidual runner, or tuned-port injection sys-tem will usually provide better performance,driveability, and fuel economy.

Tunnel-Ram Manifold —This intakemanifold is a single-plane induction sys-tem designed to produce optimum poweron all-out racing engines. The advantagesof the tunnel ram come from its combina-tion of a large common plenum and short,straight, large-volume runners. The largeplenum has plenty of space to bolt on twocarburetors, potentially flowing up to3000+cfm to optimize charge density. Thelarge plenum also minimizes pressure-waveinteraction and fuel distribution issues. Theshort runners can be kept cooler than theirlay-flat, single- and dual-plane counterparts,and they offer a straight path into the ports,optimizing ram-tuning effects.

Applications for the tunnel ram are quitelimited because of its large size; vehiclesusing tunnel-ram manifolds often require ahole in the hood and/or a hood scoop formanifold and carburetor clearance. Whilea protruding induction system may be a“sexy” addition to a street rod, in single-carburetor configurations the tunnel ramoffers very little potential power over a well-designed, single-plane manifold. Only at

very high engine speeds, with multiplecarburetors, will the advantages in the tun-nel ram contribute substantially to power.

This tunnel-ram selection can also ac-curately model fuel-injection systems withlarge, individual stacks. Strictly speaking,while the simulation combines short run-ners and a large-volume plenum, this de-sign mimics short injector stacks that opento the atmosphere quite well. For one-bar-rel-per-cylinder Weber carburetion or small-diameter, individual-injector systems, usethe Individual Runner manifold describednext. However, for large-diameter injectors,like Hillborn or Crower systems, the tun-nel-ram manifold—along with the appropri-ate airflow selection (for all cylinders com-bined)—is a good induction model.

The tunnel ram manifold selection hasthe potential to produce the highest peakhorsepower of all the manifolds listed inthe Induction menu. The large cross-sec-tional areas, straight runners, and shorttuned lengths make this manifold a “nocompromise” racing design.

Individual Runner —A manifold thatconnects each cylinder to one “barrel” ofsingle or multiple carburetors with no inter-connecting passages for shared flow is con-sidered an individual (or isolated) runnersystem (I.R. for short). A multiple Weberor Mikuni carburetor setup is a well-knownexample of this type of induction system.

This Weiand BB Chevytunnel ram manifold is asingle-plane inductionsystem designed toproduce optimum poweron all-out racing engines.It has a large commonplenum and short,straight, large-volumerunners. The tunnel rammanifold menu selectionhas the potential toproduce the highest peakhorsepower of all themanifolds listed in theInduction menu.


On a V8 engine, four twin-barrel We-bers make a very impressive sight, and atfirst glance they may appear to offer moreairflow potential that any engine needs, par-ticularly any street engine. While it maylook like overkill, the one-barrel-per-cylin-der arrangement often has substantialhorsepower limitations due to airflow re-striction! A typical Weber 48IDA carburetorflows about 330cfm per barrel. While thesum total of all eight barrels is over2600cfm (a flow rating equivalent to twoHolley Dominators), the important differ-ence here is that each cylinder can drawfrom only one 330cfm barrel. In a single-or even a dual-plane manifold, each cylin-der has access to more than one carbure-

tor barrel, reducing restriction during peakflow and increasing high-speed horse-power. While an I.R. system offers sub-stantial low-end performance benefits (moreon that next), at 5000rpm and higher ontypical smallblock installations, power canfall below the levels of an average singlefour-barrel, 780cfm induction system.

Again, on first impression, a multiple-carburetor I.R. induction may seem to of-fer way too much flow capacity, making iteasy to believe that it’s plagued with low-speed carburetion problems. Surprisingly,the same one-barrel-per-cylinder arrange-ment that produces a restriction at highengine speeds, transmits strong pressurewaves to each carburetor barrel at low

A manifold that con-nects each cylinder to asingle carburetor barrelwith no interconnectingpassages for sharedflow is considered anindividual (or isolated)runner system (I.R. forshort). Multiple Weberor Mikuni carburetorsystems are well-knownexamples of this type ofinduction system. ThisI.R. manifold wasdesigned for early OHCPontiacs.

A multiple-carburetor I.R.induction may seem to

offer way too much flowcapacity. Surprisingly,

the one-barrel-per-cylinder transmits strong

pressure waves to eachcarburetor barrel, pro-

ducing excellent throttleresponse. Individual-

runner manifolds are anoutstanding induction

choice for high-perfor-mance street engines.


speeds, producing ideal conditions for ac-curate fuel metering. Furthermore, the pres-sure waves moving in the runners are notdissipated within a plenum and don’t inter-act with other cylinders. This ensures thatthe reflected waves strongly assist cylin-der filling and reduce reversion. The com-bination of these effects makes individual-runner manifolds an outstanding inductionchoice for low-speed to medium/high-speedengine applications, such as high-perfor-mance street engines. Unfortunately, thehigh cost of these systems—and currentsmog regulations—prevents their wideracceptance.

The simulation model for the IndividualRunner choice in the Induction menu as-sumes that the runner sizes and the car-buretor venturi diameters are of “medium”dimensions. Runner length, that is, thedistance from the valve head to the top ofthe carburetors, is also assumed to be “mid-length,” and so the simulation uses a mid-range rpm power bias. These assumptionswork well with most I.R. applications, sincethis induction system is commonly usedon street engines or in road racing appli-cations that require good throttle responseand a wide power band.

The I.R. menu selection can also modelfuel-injection systems with small-diameter,medium-to-long length individual stacks.For large-diameter, short-length injectors,like drag-racing Hillborn or Crower systems,

the tunnel-ram manifold selection providesa better induction model (see the abovetunnel-ram description).

Tuned-Port Injection— This manifoldwas introduced by automakers in the mid1980’s and millions of them remain on theroad today. It represents the first mass-produced induction system that clearly in-corporated modern wave-dynamic princi-pals. To optimize low-speed torque and fuelefficiency, the TPI manifold has very longrunners (many configurations measure upto 24-inches from valve head to airbox).The runners on most TPI systems are alsoquite small in diameter—again, to optimizelow-speed power—and, unfortunately, cre-ate considerable restriction at higher en-gine speeds. Characteristic power curvesfrom this type of manifold fall slightly tosignificantly above a dual-plane up to about5000rpm, then runner restriction and anout-of-tune condition substantially lowerspeak power.

The TPI is a single-plane design thatfunctions like a long-runner tunnel ram.Each runner is completely isolated until itreaches the central plenum. This designtends to maximize pressure-wave tuningand minimize wave interactions. Since fuelis injected near the valve, the TPI systemdelivers precise air/fuel ratios with no fueldistribution or puddling problems.

There is a wide range of aftermarket

The TPI manifold wasintroduced byautomakers in the mid1980’s and millions ofthem remain on the roadtoday. It represents thefirst mass-producedinduction system thatclearly incorporatedmodern wave-dynamicprincipals.


parts available for the TPI, including en-larged and/or Siamesed runners, improvedmanifold bases, high-flow throttle bodies,and sensor and electronic modifications.The Tuned-Port Injection selection in theInduction menu models a stock TPI. How-ever, increasing the airflow (from the In-duction Airflow menu) makes it possible tomodel some of the benefits of larger run-ners and high-flow throttle bodies.

There are now many “TPI-like” and EFI(electronic fuel injection) systems availablefor small- and big-block engines. Some ofthese custom packages are based on ashort-runner tunnel ram model. Do not usethe TPI manifold model to simulate these

Some custom high-performance TPI and EFI(electronic fuel injection)

packages are based onshort-runner, high-flow

tunnel ram bases. Evensome long-runnersystems, like this

manifold from InductionTechnology, allow muchgreater airflow than the

original factory TPI.Model these manifolds

by selecting dual-plane(for small-runner sys-

tems) or tunnel ram (forlarge-runner packages)

to obtain more accuratepower curves.

The Exhaust menuselection establishes anexhaust manifold orheader configuration forthe simulated testengine. The menuincludes seven choices,four of which includemufflers. Since theprogram simulates anengine mounted on adyno, the exhaustsystem for muffledengines ends at theoutlet of the muffler.

manifolds, instead, select a single-plane (forsmall-runner systems) or tunnel ram (forlarge-runner packages) to obtain more re-alistic power curves. Only choose a TPImanifold when the induction system usesa typical long-runner TPI configuration.

THE EXHAUST MENU

The Exhaust menu, the fifth componentmenu in the main menu bar, establishesan exhaust manifold or header configura-tion for the simulated test engine. The menuincludes seven selections, four of whichinclude mufflers. Since the program isdesigned to simulate the power levels for


an engine mounted on a dyno testing fix-ture, the exhaust system for muffled en-gines ends at the outlet of the muffler anddoes not include any additional tubing usedto route exhaust gasses to the rear of avehicle.

Each of these exhaust system selec-tions apply a unique tuning model withinthe simulation. However, before we uncoverthe particulars of each choice, a short re-view of the wave dynamics acting withinthe exhaust system will help explain how“simple tubing” can boost power and en-gine efficiency. (Refer to the completeDeskTop Dynos book for a more rigorouslook at the theory of exhaust-system tun-ing.)

Wave DynamicsIn The Exhaust System

First of all, let’s begin our discussion ofexhaust wave-dynamic theory by describ-ing its antithesis: the “Kadenacy” hypoth-esis. This theory claimed that a high pres-sure “slug” of gas blasted out of the portand down the header pipe when the ex-haust valve opened. This moving mass wassaid to create a low pressure behind it,drawing additional gasses from the cylin-der. This theory is analogous to compres-sion waves traveling through a Slinky™coil-spring toy; a tight group of coils repre-senting high pressure waves moves alongthe spring followed by a more open group

of coils that represent low-pressure waves.Despite the fact that this theory was con-clusively proven to be incorrect in 1940, itis still believed by some engine “experts”to this day!

Earlier we described the high-pressurewaves that move inside the induction sys-tem. These finite-amplitude waves containso much energy that they no longer obeythe simple laws of acoustic theory. Insteadthey follow an entirely different set of rulesthat describe interactions and wave transi-tions that do not occur with low-powersound waves. Induction system energiesare just high enough to create finite-ampli-tude waves. However, exhaust systempressures are much higher and generatefinite-amplitude waves even at lower en-gine speeds.

The same laws that describe the move-ment of finite-amplitude waves and gasparticles in the induction system apply toexhaust flow. Without exploring the “depths”of finite-amplitude theory, the followingprinciples should be considered essentialknowledge to understanding wave actionin IC engine passages, particularly theexhaust system: 1) Pressure waves andgas particles do not necessarily move atthe same speeds. This phenomenon isvisible as the waves on the surface of alake wash through floating logs, pushingthem nearer the shore. The waves movefairly quickly and the logs (an analogy forthe gas particles) move more much slowly.

In an effort to explaingas flow in the exhaustsystem, it was believed

that when the exhaustvalve opened, a high

pressure “slug” of gasblasted out of the port

and down the headerpipe. As this slug

moved, it created a low-pressure “wave” behind

it, similar to the way acompression wave

travels through a

Incorrect “Kadenacy” TheoryFor Exhaust Gas Dynamics

High Pressure “Slug”(similar to compressed spring)

Low Pressure Follows(similar to expanded spring)

Slinky™ coil-spring toy. While this easy-to-visualize theory seems to make sense, itwas conclusively disproved nearly 60 years ago. Despite this, the Kadenacy effect is

still widely believed by engine “experts” to this day!


2) When a positive pressure wave reachesan area of transition, such as the end ofan open pipe, a negative suction wave iscreated that moves back up the pipe. 3)Positive pressure waves move gas particlesin the same direction as the waves; nega-tive pressure waves move gas particles inthe opposite direction of the waves.

Now let’s apply this tuning theory to theexhaust system. When the exhaust valveopens, a high pressure blast (in otherwords, a finite-pressure wave) moves downthe port and into the exhaust header pipe.This high intensity pressure wave drivesgas particles in the same direction as wavemotion and therefore assists the outflow ofexhaust gasses. When the pressure wavereaches the end of the header pipe that’sopen to the atmosphere, a negative pres-sure wave of almost the same intensity iscreated and begins to travel back up thepipe toward the cylinder. Negative pres-sure waves move gas particles in the op-posite direction as the wave, so this re-turning wave also assists exhaust gasoutflow. When the negative pressure wave(or suction wave) reaches the cylinder, itdelivers a substantial drop in pressure. Ifthis wave arrives during the valve overlapperiod (when both the exhaust and intakevalves are open) the pressure drop will help

Particle flow direction

Compression Wave Motion

Compression waves before reachingend of open pipe drive particles

in rightward direction

Returning expansion waves driveparticles in same rightward

direction

+Pa-

Particle flow direction

Expansion Wave Motion+Pa-

In order to understand howpressure waves move within

the exhaust system, theseare the two most important

things to remember: 1)When a positive pressure

wave reaches the end of anopen pipe, a negative

suction wave is created thatmoves back up the pipe and

vice versa, and 2) positivepressure waves move gas

particles in the samedirection as the waves, and

negative pressure wavesmove gas particles in theopposite direction of the

waves.

draw in fresh charge, a phenomenon calledscavenging. The overall effects substan-tially boost power by 1) assisting exhaustgas outflow, 2) beginning air/fuel chargeflow into the cylinder, and 3) helping topurge the cylinder of residual exhaust gas-ses.

Many factors contribute to optimizingthese tuning effects. Two of the most im-portant are header tubing size and length.Tubing size is really a measure of systemvolume and, on open headers at least,determines the restriction of internal pas-sages. Large, free-flowing tubes producelower pressures. Lower positive wave pres-sures generate lower amplitude suctionwaves, reducing scavenging and cylinderfilling. On the other hand, smaller diametertubing creates higher pressures that, whilegenerating strong scavenging waves, in-crease restriction and pumping work. As isthe case with every component categoryin the IC engine, the best power is pro-duced by finding a balance between twoor more counterbalancing factors. Here anoptimum balance lies between the excel-lent scavenging effects of small tube head-ers vs. the reduction in restriction andpumping work produced by large tubes.The balance tilts one way or the otherdepending on cam timing, engine displace-


ment, rpm, and several other factors!Tubing length affects when the nega-

tive suction wave arrives at the cylinder.Longer tubes delay the arrival of the scav-enging wave, appropriate for lower-speedapplications when more time elapses be-tween the opening of the exhaust valveand the overlap period. Shorter tubes re-turn a scavenging wave more quickly and“tune” at higher rpm. Once again, severalcounterbalancing factors must be includedin the analysis. The first is cam timing. Asengine speed increases, there is less timefor exhaust gasses to “blowdown” from thecylinder after the exhaust valve opens. Thehigher residual cylinder pressures increasepumping work on the exhaust stroke. Tocounteract this, earlier EVO timing willassist blowdown, but it also tends to “waste”cylinder pressures that would otherwisedrive the piston and generate power. Athigher engine speeds, the net results showbenefits from a shift toward earlier EVOtiming. To adjust for this, exhaust tubinglength must be increased to maintain thesame scavenging wave arrival time. Soagain, another pair of counterbalancingfactors are at work: Higher engine speeds

require shorter tubing lengths, but higherspeeds also require earlier EVO timing andthat needs longer header tubes to main-tain optimum scavenging. Peak enginepower at any particular rpm is produced bya balance between these factors.

There are still more factors that affectoptimum header lengths. If the primaryheader tubes terminate directly into theatmosphere, they generate a strong butnarrow suction wave. The returning waveis so “peaky” that it only assists cylinderfilling through a very narrow rpm range. Tobroaden the suction wave and extend theeffective tuning range of the exhaust sys-tem, most headers group the primary tubestogether in a larger “collector” before theyopen to the atmosphere. The collector pro-duces a transition to a larger volume be-fore the final transition to the atmosphere.This splits up or extends the width of thereturning suction wave, broadening theeffective rpm range during which theheader system can deliver an effectivescavenging wave. The diameter of thecollector dictates which end of the suctionwave has “emphasis.” Larger collectorsmimic a more direct opening to the atmo-

Negative Scavenging Wave ShouldArrive During The Valve Overlap Period

IntakeFlow

Both Intake And Exhaust Valves OpenDuring Overlap Period

ScavengingWaveScavengingWave

Exhaust tubing lengthaffects when the nega-tive suction wave arrivesat the cylinder. Longertubes delay the arrival ofthe scavenging wave,appropriate for lower-speed applications.Shorter tubes return ascavenging wave morequickly and “tune” athigher rpm.


sphere, so they emphasize the initial por-tion of returning suction wave. Smallercollectors provide less of a transition fromthe individual primary tubes and tend toemphasize the trailing edge of the suctionwave. The length of the collector changesthe time between the leading and trailingedges of the suction wave and can alsoaffect the optimum primary tube length.

Most header systems are designed sothat the lengths of the primary tubes arenearly equal. This, too, leads to anotherbalancing act between the higher numberof bends needed to obtain equal-lengthprimary tubes vs. the use of low-restrictionstraight tubes. While increased restrictionalmost always hurts performance, unequalprimary lengths can broaden the powerrange and provide more usable power inracing situations. As a result, the mosteffective header designs use as few bendsas possible to minimize restriction and re-linquish equal primary lengths to a role oflesser importance.

Exhaust Menu Selections

Now that we have peeked into the wavedynamics at work inside the exhaust sys-tem, we can turn our attention to how these

effects are simulated by the various choicesin the Exhaust component menu. As youhave discovered, the exhaust system—perhaps more than any other single part ofthe IC engine—is a virtual “playground” forfinite-amplitude waves. You are also wellaware that these interactions can be solvedonly by sophisticated, computationally-in-tensive methods that are not part of theMotion Filling And Emptying program. Whileflow restriction (back pressure) is accuratelymodeled using “pressure-drop” techniques,the effects of changes in tubing lengthsand diameters that influence the flow ofexhaust and induction gasses are closelytied to high-pressure wave dynamics. How-ever, the program does use a powerful“mini-wave model” that accurately simulatesscavenging effects for three classes ofheaders with optimum tubing lengths anddiameters. So while the program does notresolve specific header dimensions, themodel can predict engine power changesfrom various exhaust manifolds and head-ers of large and small tubing diameters(relative to the engine under test).

Stock Manifolds And Mufflers —Thefirst choice in the Exhaust menu simulatesthe most restrictive exhaust system. It as-

LESS STRONG,BUT LONGERRETURNING

WAVE

STRONG,SHORT

RETURNINGWAVE

WITH COLLECTOR

NO COLLECTOR

When primary headertubes terminate directlyinto the atmosphere, theygenerate a strong butnarrow suction wave. Thereturning wave is“peaky” and only assistscylinder filling through avery narrow rpm range.Primary tubes terminat-ing in a larger “collector”produce a wider suctionwave, broadening theeffective tuning range.


The H.P. Manifolds And Mufflers choicesimulates high-performance exhaustmanifolds designed to improve exhaustgas flow and reduce system restriction.They are usually a “ram-horn” or other“sweeping” design with fewer sharp turnsand larger internal passages, like theearly Corvette manifolds pictured above.The connecting pipes to the mufflers arelarge diameter and the mufflers generateless back pressure. This PowerPack sys-tem from Banks Engineering for 1982 to1991 Camaros/Firebirds is an excellentexample of the low-restriction systemmodeled by the H.P. Manifolds And Muf-flers choice.

sumes that the exhaust manifolds are atypical, production, cast-iron, “log-type”design, where all ports connect at nearlyright angles to a common “log” passage.These manifolds are designed more toprovide clearance for various chassis andengine components than to optimize ex-haust flow. Exhaust manifolds of this typehave widespread application on low-per-formance production engines.

The Stock Manifolds And Mufflers se-lection assumes that the exhaust manifoldsare connected to twin mufflers with shortsections of pipe. Because the engine envi-ronment is a simulated dyno cell, the ex-haust system terminates at the muffleroutlets.

The exhaust manifolds and mufflerscancel all scavenging effects, and the sys-tem is a completely “non-tuned” design. Anysuction waves that might be generated arefully damped or never reach the cylindersduring valve overlap. The restriction cre-ated by this system mimics most factorymuffler and/or catalytic converter withmuffler combinations. Back pressure lev-els in the exhaust system nearly cancelthe blowdown effects of early EVO timingand increase the pumping work losses dur-ing the exhaust cycle.

H.P. Manifolds And Mufflers —Thischoice offers a measurable improvementover the stock exhaust system modeled inthe previous selection. The high-perfor-

The Stock Manifolds AndMufflers selection as-sumes that the exhaustmanifolds are a typical,production, cast-iron “log-type” design, where allports connect at nearlyright angles to a common“log” passage.


mance exhaust manifolds simulated hereare designed to improve exhaust gas flowand reduce system restriction. They areusually a “ram-horn” or other “sweeping”design with fewer sharp turns and largerinternal passages. The connecting pipesto the mufflers are large diameter and themufflers generate less back pressure and,typically, produce more noise.

While this system is a “high-perfor-mance” design, it offers no tuning effectsand all suction waves are fully damped ornever reach the cylinders during valve over-lap. All performance benefits from thisselection are due to a decrease in pas-sage restrictions and lower system backpressure. System pressure levels mimicfactory high-performance mufflers and/orcatalytic-converter with muffler combina-tions. This exhaust system may allow somebenefits from early-EVO timing blowdowneffects (depending on the engine compo-

Here are excellent examples of high-performance “manifolds” and free-flowingmufflers from Hooker Headers. The low-restriction manifolds fit 1992-1995 Cor-vettes with an LT1 engine. The Maximum-Flow mufflers are available in 2- to 3-inchinlet/outlet configurations for Ford andChevy applications. Model these compo-nents using the H.P. Manifolds AndMufflers menu choice.

nent combination) and overall pumpingwork losses are slightly reduced by lowerback pressures.IMPORTANT NOTE FOR ALL HEADERCHOICES: Some engines, in particular, 4-or 2-cylinder applications, can develop a“full resonance” in the exhaust system—refer to the previous discussion of dual-plane manifolds for information about “full”induction system resonance. This phenom-enon can derive scavenging benefits (al-though some studies have revealed thatthe benefits are relatively small) from suc-tion waves created in the collector by ad-jacent cylinders. These “one-cylinder-scav-enges-another” tuning techniques are notmodeled in the simulation. Instead, theheaders are assumed to deliver a scav-enging wave only to the cylinder that gen-erated the initial pressure wave.

Small Tube Headers With Mufflers —This is the first component selection thatbegins to harness the tuning potential ofwave dynamics in the exhaust system.These simulated headers have primarytubes that individually connect each ex-haust port to a common collector. Thecollector—or collectors, depending on thenumber of cylinders—terminates into ahigh-performance muffler(s). Suction wavesare created in the collector, but are some-what damped by the attached muffler.Since exact tubing lengths are not simu-lated, the program assumes that the pri-mary tube will deliver the scavenging waveto the cylinder during the valve overlapperiod.

The primary tubes modeled by this menuselection are considered “small,” andshould be interpreted to fall within a rangeof dimensions that are commonly associ-


ated with applications requiring optimumpower levels at or below peak-torque en-gine speeds. These headers typically showoptimum benefits on smaller displacementengines (such as “smallblocks”), and mayproduce less power on large displacementengines. The following rules of thumb giveapproximations of tubing diameters usedby the simulation: Headers with tubes thatmeasure 95% to 105% of the exhaust-valvediameter are considered “small” for any par-ticular engine; tubes that measure 120%to 140% of the exhaust-valve diameter are“large” tube headers.

Small-Tube Headers Open Exhaust —This menu selection simulates headers with“small” primary tubes individually connect-

While not a “true” header,this tubular exhaust

system from Edelbrock forlate model cars and trucks

is CARB certified andoffers some controlledwave dynamics for im-

proved scavenging. TheFilling-And-Emptying

simulation will model theseheaders more accurately

with mufflers than without.

ing each exhaust port to a common collec-tor. The collector—or collectors, depend-ing on the number of cylinders—terminatesinto the atmosphere. Strong suction wavesare created in the collector that provide asubstantial boost to cylinder filling andexhaust gas outflow. Since exact tubinglengths are not simulated, the programassumes that the primary tube will deliverthe scavenging wave to the cylinder duringthe valve overlap period.

The primary tubes modeled by this menuselection are considered “small,” andshould be interpreted to fall within a rangeof dimensions that are commonly associ-ated with applications requiring optimumpower levels at or slightly above peak-torque engine speeds. These headers typi-

Typical small-tube head-ers are usually designed

with high-performancestreet applications in

mind. The better pieceshave 2-1/2-inch collectors

and 1-1/2- or 1-5/8-inchprimary tubes. They aremade from heavy-gauge

tubing that will withstandyears of use, and do notnecessarily have equal-

length tubes.


cally show benefits on smaller displace-ment engines but may produce less poweron large-displacement big-block engines.The following rules of thumb should give areasonable approximation of tubing diam-eters used in the simulation: Headers withtubes that measure 95% to 105% of theexhaust-valve diameter are considered“small” for any particular engine; tubes thatmeasure 120% to 140% of the exhaust-valve diameter are “large” tube headers.

Large-Tube Headers With Mufflers —This menu selection simulates headers with“large” primary tubes individually connect-ing each exhaust port to a common collec-tor. The collector—or collectors, depend-ing on the number of cylinders—terminatesinto a high-performance muffler(s). Suctionwaves are created in the collector, but aresomewhat damped by the attached muf-fler. Since exact tubing lengths are notsimulated, the program assumes that theprimary tube will deliver the scavengingwave to the cylinder during the valve over-lap period.

The primary tubes modeled by this menuselection are considered “large,” and shouldbe interpreted to fall within a range of di-mensions that are commonly associatedwith applications requiring optimum powerat peak engine speeds. These headers

Large-tube steppedheaders have large-

diameter primary tubeswith several transitionsto slightly larger tubing

diameters. These “steps”can reduce pumping

work and improvehorsepower on largedisplacement and/or

high-rpm applications.These Hooker Pro-StockBB Chevy headers have2-3/8-inch primary tubes

that step to 2-1/2-inch bythe time they reach the

4-1/2-inch collectors.

typically show benefits on high-rpm racingsmallblocks or large displacement big-blockengines. These headers may produce lesspower on small-displacement engines op-erating in the lower rpm ranges. The fol-lowing rules of thumb should give a rea-sonable approximation of tubing diametersused in the simulation: Headers with tubesthat measure 95% to 105% of the exhaust-valve diameter are considered “small” forany particular engine; tubes that measure120% to 140% of the exhaust-valve diam-eter are “large” tube headers.

Large-Tube Headers Open Exhaust —This menu selection simulates headers with“large” primary tubes individually connect-ing each exhaust port to a common collec-tor. The collector—or collectors, depend-ing on the number of cylinders—terminatesinto the atmosphere. Strong suction wavesare created in the collector that provide asubstantial boost to cylinder filling andexhaust gas outflow. Since exact tubinglengths are not simulated, the programassumes that the primary tube will deliverthe scavenging wave to the cylinder duringthe valve overlap period.

The primary tubes modeled by this menuselection are considered “large,” and shouldbe interpreted to fall within a range of di-mensions that are commonly associated


with applications requiring optimum powerat peak engine speeds. These headerstypically show benefits on high-rpm racingsmallblocks or large displacement big-blockengines. These headers produce lesspower on small-displacement engines, par-ticularly those operating in the lower rpmranges. The following rules of thumb shouldgive a reasonable approximation of tubingdiameters used by the simulation: Head-ers with tubes that measure 95% to 105%of the exhaust-valve diameter are consid-ered “small”; tubes that measure 120% to140% of the exhaust-valve diameter are“large” tube headers.

Large Stepped-Tube Race Headers —This menu selection simulates headers with“large” primary tubes individually connect-ing each exhaust port to a common collec-tor. Each primary tube has several transi-tions to slightly larger tubing diameters asit progresses towards the collector. These“steps” can reduce pumping work and im-prove horsepower as described below. Thecollector—or collectors, depending on thenumber of cylinders—terminates into theatmosphere. Strong suction waves are cre-ated in the collector that provide a sub-stantial boost to cylinder filling and exhaustgas outflow. Since exact tubing lengths arenot simulated, the program assumes thatthe primary tube will deliver the scaveng-ing wave to the cylinder during the valve

overlap period.The “stepped” design of the primary

tubes can reduce pumping work on someengines. As the high-pressure compressionwave leaves the port and encounters a stepin the primary tube, it returns a short-dura-tion rarefaction wave. This low-pressure“pulse” moves back up the header andassists the outflow of exhaust gasses.When the rarefaction wave reaches theopen exhaust valve, it helps depressurizethe cylinder and lower pumping work. Thiscan generate a measurable increase inhorsepower on large displacement and/orhigh-rpm engines.

The primary tubes modeled by this menuselection are considered “large,” and shouldbe interpreted to fall within a range of di-mensions that are commonly associatedwith applications requiring optimum powerat peak engine speeds. The following rulesof thumb should give a reasonable approxi-mation of tubing diameters: Headers withtubes that measure 95% to 105% of theexhaust-valve diameter are considered“small” for any particular engine; tubes thatmeasure 120% to 140% of the exhaust-valve diameter are “large” tube headers.

The Camshaft Menu

The final component menu allows theselection of the single most important partin the IC engine: the camshaft. For many

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STEPPED HEADERS

“Stepped” headers canreduce pumping work in

some engines. As thehigh-pressure compres-

sion wave leaves theport and encounters astep, it returns a short

duration expansionwave. These low-

pressure “pulses” moveback up the header and

assists the outflow ofexhaust gasses. This

can generate a measur-able increase in horse-power on engines that

are suffering from substantial pumping-work losses, such as large-displacement,high-rpm, drag-racing engines.


enthusiasts and even professional enginebuilders, the subtleties of cam timing defyexplanation. The reason for this confusionis understandable. The camshaft is the“brains” of the IC engine, directing thebeginning and ending of all four enginecycles. Even with a good understanding ofall engine systems, the interrelatedness ofthe IC engine can make the results of camtiming changes read like a mystery story.In many cases there are only two ways todetermine the outcome of a modification:1) run a real dyno test or 2) run a simula-tion. Since the camshaft directly affectsseveral functions at once, e.g., exhaust andintake scavenging, induction signal, flowefficiency, cylinder pressures, etc., using acomputer-based engine simulation programis often the only way to predict the out-come.

Motion Software’s Filling-And-Emptyingsimulation makes it possible to test theeffects of cam timing in seconds. The abil-ity of the program to take multiple elementsinto consideration and “add up the effectsover time” is key to analyzing the effects ofcamshaft timing changes. Unfortunately,even with such a powerful tool at hand,the subject of cam timing cannot be madesimple. The camshaft is a component thatwill thoroughly test your knowledge andcomprehension of the IC engine. Reviewthis section in light of what has been dis-cussed earlier, and we hope you’ll find adeeper understanding of camshaft opera-tion and new insight into IC engine func-

The Camshaft menu isthe final componentmenu. With it you selectcam profiles for thesimulated engine. Thefollowing pages in thischapter provide essentialinformation about camdesign and programassumptions that willhelp you use thispowerful feature ofMotion’s Filling-And-Emptying simulationsoftware.

tion.

Cam Basics

In the simplest terms, the camshaft is astraight steel or iron shaft with eccentriclobes. It is connected to the crankshaft witha chain or gear train and is usually rotatedat one-half crank speed. Lifters (or camfollowers)—and in the case of in-block camlocations, pushrods, and rockerarms—translate the rotary motion of the cam intoan up-and-down motion that opens andcloses the intake and exhaust valves. Thisentire assembly must function with highprecision and high reliability. Street enginesdriven hundreds-of-thousands of milesoperate their valvetrain components billionsof cycles. If the overall camshaft and valve-train design is good, a precision microme-ter will detect only negligible wear.

The camshaft controls the valve open-ing and closing points by the shape androtational location of the lobes. Most camsare ground to a precision well within onecrankshaft degree, ensuring that the valvesactuate exactly when intended. Timingvariations of a few degrees can develop inthe cam drive, especially in chain-drivesystems, but racing gear drives reducevariations to within one or two crank de-grees of indicated timing. Camshaft lobesalso determine how far the valves will liftoff of the valve seats by the height of thelobes (heal to toe height) and the multiply-ing ratio of the rockerarms (if used). The


Camshaft terminologycan be confusing, so

here's the low-down. Tostart off, the camshaft isa round shaft incorporat-ing cam lobes . The base

circle diameter is thesmallest diameter of thecam lobe and is shaped

perfectly round. Clear-ance ramps form the

transition from theround base circle to theacceleration ramps . Asthe cam turns, the lifter

smoothly accelerates upthe clearance ramp and

continues to rise as itapproaches the nose ,

then begins to slow to astop as it reaches

maximum lift at the lobecenterline . Maximum

Base CircleDiameter

Rotation

ClearanceRamp

Lobe Centerline

ClearanceRamp

Opening

AecelerationC

losi

ngA

ccel

erat

ion

Nose or ToeNose or Toe

Foot orFoot orHeelHeel

LifterRise

Valve-Open Duration

lifter rise is determined by the height of the toe of the cam lobe over the base circlediameter. The lifter then accelerates in the closing direction and when the valve

approaches its seat, the lifter is slowed down by closing clearance ramp. Valve-openduration is the number of crankshaft degrees that the valve or lifter is held above aspecified height by the cam lobe (usually 0.006-, 0.020-, or 0.050-inch). A symmetriclobe has the same lift curve on both the opening and closing sides; an asymmetric

lobe is shaped differently on each side of the lobe. A single-pattern cam has thesame profile on both the intake and exhaust lobes; a dual-pattern cam has different

profiles for the intake and exhaust lobes.

rates at which the valves are acceleratedopen and then returned to the seats arealso “ground into” cam lobe profiles. Onlya limited range of contours will maintainstable valve motion, particularly with high-lift, racing profiles. Unstable profiles orexcessive engine speed will force the valve-train into “valve float,” leading to rapidcomponent failure.

Visualizing And CalculatingValve Events

There are six basic cam timing eventsground into the lobes of every camshaft.These timing points are:1—Intake Valve Opening (IVO)2—Intake Valve Closing (IVC)3—Exhaust Valve Opening (EVO)4—Exhaust Valve Closing (EVC)

5—Intake Valve Lift6—Exhaust Valve LiftThese six points can be “adjusted” some-what (we’ll discuss which and how camtiming events can be altered in the nextsection), but for the most part they are fixedby the design of the cam. Other timingnumbers are often discussed, but they arealways derived from these basic six events.These derivative events are:7—Intake Duration8—Exhaust Duration9—Lobe Center Angle (LCA)10—Valve Overlap11—Int. Center Angle or Centerline (ICA)12—Exh. Center Angle or Centerline (ECA)

The first four basic timing points (1 thor-ough 4) pinpoint the “true” beginning andend of the four engine cycles. These valveopening and closing points indicate when


the function of the piston/cylinder mecha-nism changes from intake to compression,compression to power, power to exhaust,and exhaust back to intake. What could bemore important from the standpoint of un-derstanding engine function and perform-ing engine simulations?

Compared to the basic timing events,many simulation experts believe that mostof the second six timing values are notonly unimportant, they actually “blowsmoke” over the whole issue of cam timinganalysis. Naturally, four of the second sixevents are the most publicized by the cammanufacturers and enthusiast magazines.This unfortunate situation evolved overmany years of selling and marketing cam-shafts in the automotive aftermarket. Longbefore engine simulations were widely usedand a good understanding of the relation-ship between the individual valve eventsand engine power ever existed, the de-scriptions of the distance between valveevents rather than the valve events them-selves became a standard measure of acamshaft.

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60° ABDC24° ATDC24° BTDC60° BBDC

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Lobe Center Angle 108°

Crane “short event”camshaft

Exhaust closes3° before TDC

BDC

TDC

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VALVE MOTION CURVES

Since this assortment of cam specifica-tions is used almost interchangeably, it isalmost impossible to “talk camshaft” with-out a good understanding of all these terms.Probably the best way to organize andvisualize this nomenclature is to picture thecommon “twin-hump” event drawing. Thisillustration depicts valve-motion curves(sometimes called valve displacementcurves) for the exhaust lobe on the leftand the intake lobe on the right, locatingthe valve overlap period and TDC (top deadcenter) at the center of the graph. If youbecome sufficiently familiar with this draw-ing so that you can easily picture it in yourmind, you should be able to figure out anyof the cam timing specs and how they relateto one another.

The graph plots crank degrees on theX-axis (left to right) and valve lift on the Y-axis (up and down). The width of the graphis slightly shortened. Since no valve mo-tion occurs during about 200 degrees ofcrank rotation when the “true” compressionand power strokes take place, the graphchops off 60-degrees on each end, run-

The best way to visualize camshaft timing is with this “twin-hump” event drawing. Itdepicts the valve-motion curves for the exhaust lobe on the left and the intake lobeon the right, locating the valve overlap period and TDC at the center. If you becomesufficiently familiar with this drawing so that you can easily picture it in your mind,you will be able to quickly evaluate any cam timing specs and visualize how theyrelate to one another.


ning from 300-degrees before to 300-degress after TDC (total four-cycle sweepis actually 720 degrees).

Picture events on the graph taking placefrom left to right, passing through TDC atthe center. The left “hump” is the exhaustvalve motion curve. Focus your attentionat the 240-degree point before TDC. Notethat the 180-degree vertical line is actuallya BDC (bottom dead center) marker, so240-degrees before TDC lies 60-degreesbefore BDC. This is the EVO (exhaust valveopening) timing point, the beginning ofexhaust blowdown and rapid cylinder de-compression. Follow the exhaust valvecurve through its maximum lift—about0.450-inch—that occurs at 108-degreesbefore TDC to its closing point at 24-de-grees after TDC. Now notice the intakevalve motion curve. At 24-degrees beforeTDC, when the exhaust valve is still open,the intake valve leaves its seat and beginsto follow its motion curve. During the spacebetween 24-degrees before TDC and 24-degrees after TDC, both valves are open,defining the overlap period when exhaustscavenging can take place. Now continueto follow the intake motion curve to its

Lobe Center Angle is the angle mea-sured in camshaft degrees (multiply bytwo for equivalent crankshaft degrees)

between the maximum-lift points on theintake and exhaust lobes for the same

cylinder. The lobe center angle is“ground” into the cam when it is manu-

factured and cannot be changed (unlessthe cam is reground). As the lobe-center

angle is decreased, the valve overlapperiod (when both intake and exhaust

valves are open) is increased. The lobecenterline angles are the angles mea-

sured in crankshaft degrees between thepoints of maximum lift on the intake and

exhaust lobes and Top Dead Center.These values are determined by the

“indexing” of the cam to the crankshaft.

EXHAUST LOBEINTAKE LOBE

Valve Overlap

Lobe-Center Angle

Centerline Angles

EXHAUST LOBEINTAKE LOBE

TDCTDC

Valve OverlapValve Overlap

Exhaust Centerline

IntakeCenterline

maximum lift—also about 0.450-inch—at108-degrees after TDC. The intake closingpoint occurs at 240-degrees after TDC or60-degrees after BDC (the 180-degreeline). You have now traced out the six basiccam timing points. Let’s see how theserelate to the six derivative events.

First consider duration—the number ofcrank degrees that the valves are off theirseats. This “length of time,” expressed incrank degrees, is a measure of the trueintake and exhaust cycles. The exhaustvalve opens 60-degrees before BDC andcloses 24-degrees after TDC, and since180 degrees of crank rotation exists be-tween BDC and TDC, the exhaust dura-tion is: 60 + 180 + 24 = 264 degrees off-seat duration. In the case of the intakevalve, valve opening occurs 24-degreesbefore TDC and closing 60-degrees afterBDC, so the intake duration is: 24 + 180 +60 = 264 degrees. Duration is easy to cal-culate from the opening and closing points;however, watch out for the one “tricky”element: short events. For cams that openthe intake valve after TDC instead of be-fore TDC or close the exhaust valve be-fore TDC instead of after TDC (common


with 0.050-timing specs), make sure tosubtract the timing point instead of addingit to calculate the duration. For example, aCrane Cams, Inc. profile HMV-272-2-NCopens the exhaust valve 51-degrees be-fore BDC and closes it 3-degrees beforeTDC. To calculate the duration subtract the“short” closing point: 51 + 180 –3 = 228degrees. While these “short” timing eventsare much more common when working with0.050-inch timing specs, you may alsocome across seat-to-seat timing specs foremissions-restricted camshafts with “short”events, designed to minimize or eliminatevalve overlap. In a nutshell, the duration issimply the number of crank degrees sweptout by the lift curve.

Next, let’s look at the three centerlineangles: Intake Centerline, Exhaust Center-line, and Lobe Center Angle (sometimescalled Lobe Centerline). The terms “Cen-terline” and “Center Angle” are methods ofdescribing the distance to or from the ex-act center of a lobe. The Intake and Ex-haust Centerlines describe the distancefrom the center of the Intake and Exhaustlobes to TDC. Both of these timing specsare measured in crank degrees. Thatmeans that the number of degrees thecrank rotates from the point at which thepiston rests at TDC until the lifter contactsthe exact center of the intake lobe is theIntake Centerline. When the lobes are sym-metric, that is they have the same shape

The Lobe Center Angledescribes the angulardistance between the

centers of the intake andexhaust lobes as viewed

on the cam itself . Thisdistance is a measure-ment of the cam—not

relative to TDC—and isthe only cam spec that

cannot be altered regard-less of how the cam is“degreed” with crank-shaft. Because of this,

the LCA is measured incam degrees.

and valve motion rates on the opening andclosing sides, the Intake Centerline andExhaust Centerlines will occur at the pointof maximum valve lift. Asymmetric profilesare quite common, although, the amountof asymmetry typically is very small, so thecenterlines should still fall within two orthree of degrees of maximum lift point.

All but one of the basic and derivativecam-timing specs are measured in crankdegrees. This makes sense since the tim-ing specs fundamentally describe valve po-sitions (and durations) as they relate topiston positions (and piston movement).This applies to every timing spec but LobeCenter Angle (LCA). The LCA is meant todescribe the angular distance between thecenters of the intake and exhaust lobes asviewed on the cam itself. This distance isa cam-specific measurement, not relativeto TDC, and is the only cam spec thatcannot be altered regardless of how thecam is “degreed” with crankshaft. It is saidto be “ground into” the cam. Because thisspec relates one lobe position to the other,independent of the engine or crankshaft, itis measured in cam degrees. The numberof crank degrees between the center ofthe lobes is twice the LCA.

All of the camshaft specifications de-scribed thus far tell something about howthe cam is manufactured and how it shouldbe installed in the engine. To help keepthem straight in your mind, let’s reorganize


them in two new groups. The first groupcontains all the timing specs that are mea-sured from TDC or BDC piston positionsand are dependent on how the cam is in-stalled, or indexed, in the engine:1—Intake Valve Opening (IVO)2—Intake Valve Closing (IVC)3—Exhaust Valve Opening (EVO)4—Exhaust Valve Closing (EVC)5—Int. Center Angle or Centerline (ICA)6—Exh. Center Angle or Centerline (ECA)Each of these timing specs indicate a open-ing, closing, or lobe centerline point mea-sured from a TDC or BDC piston position.If the cam is installed in an advanced orretarded position relative to TDC, the valueof these timing specs will change. In fact,advancing and retarding the cam is oneway to change cam timing that we’ll dis-cuss in detail later in this guide.

The following group of cam timing events

and specs at dependant on TDC or crank-shaft position:7—Intake Duration8—Exhaust Duration9—Lobe Center Angle (LCA)10—Valve Overlap11—Intake Valve Lift12—Exhaust Valve LiftEach of these six terms compare one camspec to another; they are not measuredrelative to any fixed crank position. Referagain to the valve-motion plot to confirmthis. Remember that the duration of eachlobe is the distance in crank degrees be-tween the valve opening and closing points.Sliding these curves left or right (analo-gous to advancing or retarding the cam)does not change the distance between theopening and closing points and does notchange duration. Duration is, therefore,another “ground in” cam spec. The same

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IVO IVCEVO EVC

ICAECAVALVE MOTION CURVES

ExhaustCenterline

IntakeCenterline

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264 DEGREES INTAKE DURATION

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108° LobeCenter Angle

48° VALVE OVERLAP

EXHAUSTLIFT

INTAKELIFT

BDC

TDC

BDC

VALVE MOTION CURVES

This group of valveevents are all

measured from TDCor BDC piston

positions and aredependent on how

the cam is installed,or indexed, in the

engine.

These cam timingevents compareone cam spec toanother and arenot dependant onTDC or crankposition.


limitation applies to valve overlap. Thelobes are spaced by a fixed LCA, and theycan’t move with respect to each other, sothe length of the overlap period—the dis-tance between EVC and IVO at a specificlifter rise—never changes for a particularcamshaft. Finally, valve lift is another com-parison of cam specs. In this case, it’s thedistance between the lobe heal height andtoe height multiplied by the rocker ratio.Valve lift is measured in inches (or milli-meters) and is never related to crank angleor TDC.

Take some time to review the relation-ships between all of the cam specs andthe valve motion drawings. Trace the ac-tion of the valves from left to right throughthe exhaust cycle and, as TDC approaches,through overlap, then continue to the rightthrough the intake cycle. These drawingsgive an excellent mental “picture” of therelationships between LCA and the indi-vidual ICA and ECA values. In fact, as-suming a symmetric profile, it is possibleto calculate all of the center angles, dura-tions, and the overlap from the four basicvalve events (EVO, EVC, IVO, IVC). It isalso possible to calculate the four valveevents from the duration, the LCA, andeither the ICA or the ECA. We’ll discussthe details of converting timing specifica-tions using information from cammanufacturer’s catalogs later in this guide.

How Valve-Event TimingAffects Power

As we mentioned earlier, years of mar-keting efforts by cam manufacturers haveestablished an accepted methodology for

identifying and classifying camshafts. Thespecifications most commonly listed bymanufacturers are:1—Intake Duration2—Intake Valve Lift3—Exhaust Duration4—Exhaust Valve Lift5—Lobe Center Angle (LCA)6—Intake Center Angle or Centerline (ICA)

Long before engine simulations werewidely used and designers gained an un-derstanding of how the changes in valve-event timing affect power, “manufacturer’scatalog” specs became a standard mea-sure of cam profiles. Unfortunately, theseterms place the emphasis on the spanbetween the valve events rather than onthe events themselves. For example, it iscommon to compare two cams by compar-ing their intake and/or exhaust valve-openduration. While duration does point to theintended use for the cam, it doesn’t indi-cate the valve events, making it difficult topredict engine performance. If Cam A has264 degrees of exhaust duration and CamB has 300 degrees, the longer durationspec doesn’t give a clue about how theadditional valve-open timing will be allo-cated to the opening and closing events.Are the entire forty degrees added to CamB’s valve opening point; are they added tothe closing point; or are they split betweenthe two in some proportion? Without know-ing the exact valve events, one can onlyguess at the outcome. A more critical situ-ation exists for engine simulation programs:Without knowing the exact valve events, asimulation isn’t even possible! Not usingexact valve timing events during a simula-tion is like building and testing an engine

Before engine simula-tions were widely used,cam manufacturersestablished a methodol-ogy for identifying andclassifying camshafts.Unfortunately, these“catalog” specs placethe emphasis on thespan between the valveevents rather than on theevents themselves.


without defining when the valves open andclose; the whole concept doesn’t make anysense.

The emphasis on event timing in en-gine simulation programs has driven manyleading-edge designers to discount whatthey now term as ambiguous or less-use-ful cam timing specifications, in particular,advance and retard figures, centerlines, anddurations. For some, this may be a difficultparadigm shift, but the rewards are sub-stantial: you may find a new understand-ing of the IC engine at the end of yourefforts. The next few paragraphs delve intothe effects of individual valve eventslearned from both real-world and simulatedtesting. We’ll also relate these basic timingevents to other popular cam specifications,since it will be many years—if ever—be-fore some of the less-relevant specs dis-appear from cam manufacturer’s catalogs.

The four basic valve-event timing points(EVO, EVC, IVO, IVC) can be grouped intothree categories based on their influenceon engine performance: 1) EVC and IVOare the least important individually, but

A manufacturer’s catalog lists Cam A as having 264 degrees of duration while Cam Bhas 300 degrees. But just the duration spec doesn’t give a clue about how the addi-tional valve-open timing will be allocated to the opening and closing events. Thesevalve motion curves indicate just one of an infinite number of possibilities. Withoutknowing the exact valve events, running an engine simulation isn’t possible.

Cam B

Cam A Cam A

Cam B

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INTAKE EXHAUST

BDC

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BDC

VALVE MOTION CURVES

together comprise the overlap period thathas a significant effect on power, and 2)the EVO is the next most important timingpoint since is determines the beginning ofthe exhaust cycle and cylinder blowdown,and 3) IVC is the most critical since it fixesthe balance between cylinder filling andintake reversion, each having a potent ef-fect on engine output.

EVC/IVO, The Valve Overlap Period —The valve overlap period occurs as thepiston passes through TDC after the mainportion of the exhaust stroke. The intakevalve opens before TDC (usually) and sig-nals the beginning of period of time duringwhich both intake and exhaust valves areoff their seats. As we found in our discus-sion of exhaust systems, a high pressurewave produced when the exhaust valveopens (EVO) travels to the end of theheader and returns a strong negative pres-sure wave that delivers a pressure drop atthe exhaust valve. If this pressure droparrives during the overlap period, it will helppurge the cylinder of exhaust gasses and,


despite the upward movement of the pis-ton, begin the inflow of fresh charge fromthe induction system. This phenomenon iscalled scavenging, and an engine thatdelivers its scavenging wave during theoverlap period is said to be “in tune.” If thepressure wave is early or late, or not evencreated by exhaust system, the piston ris-ing in the bore during the first part of over-lap will force exhaust gasses into the in-duction system, producing a phenomenoncalled “reversion.” When this occurs, camtiming and the exhaust system are “out oftune.”

Reversion is a power killer. When ex-haust gasses are driven into the inductionsystem, they force air/fuel mixtures backupstream. Severe reversion can drive theair/fuel charge out of the air inlet, creatinga “standoff” of vapors above the carbure-tor. When the charge is drawn back intothe engine, it is re-atomized with fuel cre-ating “double-rich” mixtures. Additional, less

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EXPANSION WAVE COVERSENTIRE VALVE OVERLAP PERIOD

EXHAUSTLIFT

INTAKELIFT

BDC

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BDC

Expansion Wave MovingIn Header Pipe Represents Port/Valve

Boundry

+Pa-

The goal of the engine designer is to “cover” the overlap period with the arrivingscavenging wave during as wide an rpm range as possible. When this is done, theexhaust system and the engine are said to be “in tune” during that range of enginespeeds. If any part of overlap does not coincide with the presence of a low-pressurescavenging wave, reversion or reduced cylinder filling will drive the engine partiallyout of tune.

severe, symptoms of reversion include adrop in manifold vacuum, rough idle and/or high idle speeds, and a substantial in-crease in emissions from over-rich mix-tures.

An engine that develops reversion andruns poorly at lower engine speeds mayrun fine at higher speeds. In fact, mostrace engines exhibit these symptoms; thecommon side effects of a high-speed racetune. The scavenging wave that missedthe overlap period at low speeds may re-turn on time at higher engine speeds, op-timizing tuning, cylinder filling, and poweroutput. The nature of the exhaust systemis to create in-and-out-of-tune conditionsas the engine moves through its rpm range.The goal of the engine designer is tobroaden the arrival of the scavenging waveas much as possible (remember our dis-cussion of header-pipe collectors) “cover-ing” the overlap period through as much ofthe max-power rpm range as possible. If


any part of overlap does not coincide withthe presence of a low-pressure scaveng-ing wave, reversion or reduced cylinderfilling will drive the engine partially out oftune.

Valve overlap periods vary from zero toabout 40-degrees on street and basic per-formance engines. All-out racing enginesuse overlap periods as long as 120 de-grees. Engines with long overlap tend tohave very “peaky” power bands, since areturning scavenging wave “covers” thevalve overlap period through only a rela-tively narrow rpm range. However, wideoverlap periods are needed to effectivelyscavenge large-displacement engines athigh engine speeds. Shorter overlap peri-ods, on the other hand, stay in-tune througha wider rpm range, but tend to reduce peakhorsepower because they limit scaveng-ing. The number of overlap degrees is a

good indicator of the intended use of thecam: short overlap for broad torque andgood low-speed power, long overlap forhigh power at high engine speeds.

Lobe Center Angle and Overlap —Since valve overlap is directly related tothe Lobe Center Angle (LCA)—refer againto the valve motion diagram—is has be-come commonplace to discuss changes invalve overlap as being synonymous withchanges in LCA. In fact, some believe thatchanging LCA (requires regrinding cam) isthe correct method of adjusting overlap.This is not true. This common misconcep-tion comes from an inadequate understand-ing of the function and effects of the indi-vidual valve-event timing points. For ex-ample, suppose that you wish to increasethe valve overlap period for a particularcamshaft. Decreasing the LCA will, in fact,

Decreasing the LCA will directly increase valve overlap by moving IVO earlier andEVC later. However, if the goal is to boost high-speed horsepower (the reason foradditional overlap), narrowing the LCA also moves the IVC earlier. This reduces “rameffects” in the induction system at higher engine speeds, clearly the wrong approachfor performance. In addition, a narrower LCA opens the exhaust valve (EVO) later andthat delays cylinder blowdown, an effect that tends to boost low-speed—not high-speed—power. The correct method of adjusting overlap, and every other cam timing,should be to apply the appropriate changes to the individual valve opening and clos-ing events.

IVO EarlierYES!

EVC LaterYES!

EVO LaterNO!

IVC EarlierNO!

CHANGING LCA ALONE IS A “NONSENSE” TUNING APPROACH

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BDC

VALVE MOTION CURVES


directly increase the overlap, but let’s lookat the consequences. If the goal is to in-crease overlap, the need must be to boosthigh-speed horsepower. With a smallerLCA, the EVC occurs later and IVO isearlier, both of these changes directly in-crease overlap and tend to boost highspeed power. So far, so good. However,changing the LCA has the effect of “rotat-ing” the lobes on the camshaft and movingthe IVC earlier. Since later IVC timing tendsto take advantage of the “ram effects” inthe induction system at higher enginespeeds, closing the intake valve earlier isclearly the wrong approach. In addition, witha narrower LCA, the EVO occurs later andthat delays cylinder blowdown, anothereffect that tends to boost low-speed power.Again, the wrong approach. This “schizo-phrenic” or nonsensical tuning approachapplies the correct EVC and IVO timing,but counteracts these effects with earlierIVC and later EVO timing; the net resultsare often a modest shift in tuning towardhigher engine speeds. The correct methodof adjusting overlap, and every other camtiming alteration, should be to apply theappropriate changes to the individual valveopening and closing events. In this case, alonger overlap should be produced byapplying an earlier IVO and EVO and alater IVC and EVC. Yes, this also increases

duration, but applying individual valve-eventchanges that complement—rather thatcounteract—each other, produce a muchmore effective camshaft for the intendedpurpose.

EVO, The Exhaust Valve OpeningPoint —EVO timing is the next most im-portant of the basic valve events. Changesin this timing point can have a substantialimpact on engine efficiency and perfor-mance. The opening of the exhaust valvecreates a high-pressure wave that travelsthrough the exhaust system that, with prop-erly designed headers, returns as a strongscavenging wave during the overlap pe-riod. EVO timing must be coordinated withheader tubing length, engine speed, andvalve overlap to obtain full benefit fromscavenging. The EVO point also signalsthe beginning of the blowdown phase ofthe exhaust cycle by starting the rapiddecompression of the cylinder at the endof the power stroke.

The optimum EVO timing point is an-other “moving target” for the engine de-signer. It should be no surprise that aperfectly timed EVO simply doesn’t exist.However, if perfection were possible, anideal EVO would delay the opening of theexhaust valve until the piston reached BDC,allowing the engine to harness every last

800

1000

600

400

14.7

0 180 360TDCTDC

Degrees of Crankshaft RotationBDC

EVO Point

Intake

Ignition

Compression Power Exhaust

BDC720540

Cyl

ind

er P

ress

ure

(p

si)

EVO signals the begin-ning of the blowdownphase by starting the

rapid decompression ofthe cylinder near the endof the power stroke. The

opening of the exhaustvalve also creates a

high-pressure wave thattravels through the

exhaust system that,with properly designed

headers, returns as astrong scavenging wave

during the overlapperiod.


bit of energy generated by combustion.Then, precisely at BDC, the exhaust valvewould “pop” open and, miraculously, all ofthe residual gas pressure would vanish(even better, a slight vacuum would de-velop in the cylinder). Then as the pistonmoves from BDC to TDC, no horsepowerwould be wasted on “pumping” spent gas-ses from the engine. Unfortunately, the realworld of engine gas dynamics is far fromperfection. Substantial power is consumedby driving the piston up the bore on theexhaust stroke, especially at high enginespeeds with large-displacement enginesthat generate prodigious amounts of ex-haust gas. To offset these losses, earlyEVO timing starts the blowdown of high-pressure gasses before the exhaust strokeeven begins. But the reduction in pumpingwork doesn’t come without a drawback;earlier EVO timing “wastes” some of thepower-producing gas pressure from com-bustion. The balance between these twofactors optimizes power, but the balancechanges as engine speed changes (and,of course, it also changes as displacement,flow restriction, and the timing of other valveevents change).

To complicate matters even further, oncean optimum EVO timing has been found,we need to consider how this timing af-fects the arrival of the scavenging wave.Remember, that the chain of events thatleads to the arrival of the scavenging wave

Header lengths need tobe adjusted to harnessEVO timing and scaveng-ing effects to full benefit.However, header tubingdiameter also plays animportant part in deter-mining exhaust systemrestriction and canchange blowdown andpumping-loss character-istics. Despite the factthat EVO is an extremelyimportant timing event, acombination of many

tuning factors make it very difficult to determine the outcome of specific changes toEVO timing. Motion’s Filling-And-Emptying simulation can “home-in” on the optimumcombination from a large number of interdependent conditions.

begins at the opening of the exhaust valve.When the pumping-loss/blowdown balancehas been found, header lengths (and over-lap timing) may need to be adjusted toharness scavenging effects to full benefit.Then when you consider that header tub-ing diameter plays a part in determiningexhaust system restriction and can changeblowdown and pumping-loss characteris-tics, the jumble of tunable elements growseven larger. Finally, if the induction systemis improved, higher post-combustion pres-sures can worsen the pumping problem andthat may require further changes to EVOtiming, starting the whole tuning processover again.

Despite the fact that EVO is an ex-tremely important factor in engine output,this web of interrelated effects make it verydifficult (probably impossible) for engineexperts to determine the outcome of spe-cific changes to EVO timing. Here is an-other example of how engine simulationsare useful. A series of simulations can“home-in” on the optimum combination froma large number of interdependent condi-tions.

IVC, The Intake Valve Closing Point —IVC is the most critical of the basic valvetiming events. The intake valve closingpoint establishes a balance between cylin-der filling and intake reversion, each hav-ing a potent effect on engine performance.


After the intake stroke is completed andthe piston reaches BDC, the column of air/fuel mixture moving through the inductionsystem has built up considerable momen-tum (the “ram tuning” effect). This internalenergy forces additional air and fuel to flowinto the cylinder even as the piston beginsto move up the bore on the early part ofthe compression stroke. At some point,however, the pressure in the cylinder be-gins to exceed the pressure of the incom-ing charge, and the induced charge startsto “revert” or flow back into the inductionsystem. This is the ideal point for IVC,because the cylinder has received thegreatest volume of air and fuel and willgenerate the highest pressures on thepower stroke.

Unfortunately, optimum IVC occurs onlyat one engine speed and is dictated bycylinderhead flow, induction ram-tuningeffects, intake valve opening timing, and ofcourse by engine rpm. The longer the in-take valve is held open the more “peaky”engine performance becomes. Late IVCcan create induction flow reversion with anaccompanying drop in manifold vacuum,rough idle and/or poor idle quality, and anincrease in emissions from over-rich mix-tures (rich mixtures are created when thecharge is forced back up the manifold and

��

��

OptimumIntake Valve

Closing Point

IVC is the most critical of thebasic valve timing events. Theintake valve closing pointestablishes a balance betweencylinder filling and intakereversion, each having a potenteffect on engine performance.When the pressure produced inthe cylinder as the piston beginsto move up the bore on thecompression stroke exceeds thepressure of the incoming charge,the induced charge starts to“revert” or flow back into theinduction system. This is theideal point for IVC, because thecylinder has received thegreatest volume of air and fueland will generate the highestpressures on the power stroke.

passes through the carburetor—or by thefuel injector—a second time). When theinduction system on a racing engine isproperly designed, the pressure wave cre-ated when the intake valve opens is re-turned to the cylinder as a strong suctionwave just about the time cylinder pressuresbegin to overcome the ram tuning effects.This momentary drop in pressure allows abit more cylinder filling and a slightly laterIVC. This critical tuning can add the win-ning edge to Pro Stock engines or othermax-power applications, however, induc-tion system design, IVO, and IVC timingmust all be synchronized to produce theseeffects. Since induction components aretypically hand built “one offs” on enginesof this type, custom cams are ground forindividual engines to optimize power inthese competitive classes.

On the other end of the spectrum, if thegoal is to build an engine that performswell throughout a wide rpm range and of-fers good idle characteristics, late IVC isdefinitely the wrong approach. The intakevalve must close early enough to preventreversion at lower engine speeds, an es-sential step in producing low-speed torque.However, early IVC limits cylinder filling athigher engine speeds and reduces peakpower. IVC timing in the high 50-degree


range is typical for a mild street cam, thehigh 60-degree to the mid 70-degree rangeis common on high-performance and mildracing grinds, and IVC from the 80-degreeto low 100-degree range is all-out racingtiming.

Camshaft Menu—Lifter Choices

The camshaft menu consists of twogroups. The upper category provides a listof three common lifter types used to modelcamshaft acceleration rates. The lower partof the menu offers an application-specificgroup of individual cam grinds, and the finaltwo “Other” choices allow the direct entryof cam timing events to simulate virtuallyany camshaft. Making a selection fromeach of these two groups “programs” thesimulation to develop a valve motion curvefor the selected camshaft.

The Filling-And-Emptying program usesa sophisticated model to accomplish thissimulation, but the user must keep in mindthat valve motion curves for both the in-take and exhaust valves are being simu-lated from only six data points, three forthe intake valve and three for the exhaustvalve. The starting point for each simula-tion is the opening and closing timing andthe lobe lift. From these three points, andthe lifters selection, the program creates amotion curve that pinpoints valve lift at eachdegree of crank position. The results areremarkably accurate, however, the simula-tion cannot model subtle differences be-

The upper category ofthe camshaft menu

provides a list of threecommon lifter types. The

lower part of the menuoffers an application-

specific group ofindividual cam grinds,

and the final two “Other”choices allow the direct

entry of cam timingevents to simulate

virtually any camshaft.

The Filling-And-Emptying program mod-els symmetric valve motion curves fromsix data points, three for each lobe: 1) theopening point, 2) the closing point, and3) the lobe lift. Although some cam grindsare asymmetric, performance differencesbetween a symmetric model and actualasymmetric valve motion is quite small.

Rotation

CalculatedLobeCenterline(assumingsymmetricprofile) Lifter

Rise

ClosingPoint

OpeningPoint

Lobe Profiles Are CalculatedFrom Opening Point, Closing Point,

And Lifter Rise

tween cam grinds that use the same eventtiming and valve lift specs. Furthermore,the model develops a symmetric valvemotion curve, although some cam grindsare asymmetric (meaning that the “open-ing” side of the lobe differs in shape fromthe “closing” side). Asymmetric modeling


is impossible with only three data inputpoints, luckily, performance differencesbetween symmetric models and actualasymmetric valve motions are often quitesmall.

The first part of the Camshaft menuoffers three choices: 1) Hydraulic Flat-Tap-pet Lifters, 2) Solid Flat-Tappet Lifters, and3) Roller Solid Or Hydraulic Lifters. Eachof these choices instructs the simulation toapply a unique “ramp-rate” model to thevalve motion curve. The first two choicesare flat-tappet lifters. This lifter uses a flatsurface to contact or “rub” on the camlobes. Flat-tappets are simple and quitereliable in stock and many high-perfor-mance applications. However, their designlimits the rate at which the valves can beopened and closed. Slower valve accel-eration reduces the exposed curtain areaand the flow capability of the cylinderheadsat every point during the lift curve, exceptfor the fraction of a second when the valvepasses through maximum lift. (Interestingly,flat tappet lifters can be made to out accel-erate roller lifters during the first severalhundredths of an inch of the lift curve, butroller lifters easily surpass flat tappet liftrates by the time the valves reach 0.100-to 0.200-inch of lift.) The third choice in thelifter group is a roller lifter (either solid orhydraulic, more on these differences next).This design incorporates a cylindrical ele-ment that rolls over the cam lobe. Whilethere are slight gains from reduced fric-tion, the greatest benefit from a roller cam/lifter design lies “hidden” in the mechanicalrelationship between the roller and camlobe. Roller cams can be ground with pro-files that generate very high accelerationrates, opening and closing the valves muchmore quickly than flat-tappet cams. Fastervalve acceleration increases the averagecurtain area exposed throughout the liftcurve. This can substantially improve cyl-inderhead flow and horsepower.

The three choices in the lifter menu, aspreviously indicated, establish a “ramp-rate”model for the simulated valve-motion curve.The lowest acceleration is assigned to thefirst menu choice: Hydraulic Flat-TappetLifters. Hydraulic lifters incorporate a self-adjusting design that maintains zero lash

Rotation

Retaining Ring

Pushrod Seat

Oil Inlet

Metering ValveAdjustingOilCavity

Check Valve

Lifter Body

Lifter Face

Hydraulic Lifter

Street/MildPerformance

Lifter Face

Rotation

Retaining RingPushrod Seat

Oil MeteringValve

Oil InletLifter Body

Solid Lifter

High Perf./Racing

Rotation

Retaining Clip

Pushrod Seat

Oil Inlet

Check ValveAdjustingOilCavity

Lifter Body

Roller BearingsLifter Cam Roller

Metering Valve

Roller Lifter

High/Perf.All-Out Racing

The three lifter choices establish a “ramp-rate” model for the simulated valve-mo-tion curve. The lowest acceleration is as-signed to Hydraulic Flat-Tappet Lifters .The next highest acceleration is appliedto Solid Flat-Tappet Lifters . The highestacceleration is reserved for the last menuchoice: Roller Solid Or Hydraulic Lifters .


in the valvetrain. They are well-known forproviding quiet, trouble-free operation inmild- to high-performance street engines.Hydraulic, flat-tappet cam profiles usuallygenerate low acceleration rates to optimizevalvetrain reliability and extend engine life.These are the characteristics of the modelused by the simulation program whenHydraulic Flat-Tappet Lifters is chosen. Thenext highest acceleration rates are as-signed to Solid Flat-Tappet Lifters. Theselifters incorporate no lash adjusting mecha-nism and require an operating clearance(or lash) in the valvetrain, usually 0.020- to0.030-inch. Clearance is typically adjustedat the rockerarm or with spacers in thecase of overhead cams with cam follow-ers. Solid lifter cams are often ground withfaster acceleration rate ramps, generatemore valvetrain noise and wear, and aredesigned for performance-oriented applica-tions. These more aggressive characteris-tics are used by the simulation to derive avalve-motion curve when Solid Flat-Tap-pet Lifters is chosen from the menu. Fi-nally, as described above, the highest ac-celeration rates are applied to the last menuchoice: Roller Solid Or Hydraulic Lifters.This choice applies to very aggressive rampacceleration rates and derives valve mo-tion curves appropriate for most racing,

roller-lifter camshafts.The simulation uses increasing valve-

train acceleration to model hydraulic, solid,and finally roller-lifter camshafts. This is agood assumption, since cams typically uselifters that are suited for the intended ap-plication, and cam profiles for specific ap-plications typically apply predictable valveacceleration rates. However, this is not al-ways the case. For example, some cam-shafts currently available for mild streetengines use roller lifters, not to achievehigh valve acceleration rates, but to opti-mize reliability. In these cases, choosingroller lifters from the Camshaft menu willproduce optimistic power curves from thesimulation. So, to improve program accu-racy, ask yourself if the camshaft you aremodeling fits the following application-spe-cific description before you make a lifterselection:

Menu Choice Intended ApplicationHydraulic Flat-Tappet Street/HPSolid Flat-Tappet HP/RacingRoller Very HP/Racing

If the cam you’re modeling is a roller-liftergrind but a very mild-street profile, selectHydraulic or Solid Flat-Tappets from themenu since this choice will produce a lift

To improve programaccuracy, ask yourself if

the camshaft you aremodeling fits the applica-

tion-specific descriptionlisted in the above text. If

your cam uses roller liftersbut is a mild street profile,

select Hydraulic or SolidFlat-Tappets from the

menu since these choiceswill produce a lift curve

that matches a mildcamshaft. On the other

hand, if the cam is a high-performance grind, select Solid Lifters or Roller Lifters since these will model the

faster acceleration rates of an aggressive performance grind. If you are modeling alarge-diameter, solid-lifter racing cam, like some “mushroom” lifter grinds, the Solid

Lifter choice may underestimate the acceleration rate of these competition camshafts.In this case you may find more accurate predictions from the Roller Lifter selection.


curve that best matches a mild street cam-shaft. On the other hand, if the cam is ahigh-performance grind, select Solid Lift-ers since this will model the faster accel-eration rates of aggressive performancegrinds. If you are modeling a large-diam-eter, solid-lifter racing cam, like some“mushroom” lifter grinds, the Solid Lifterchoice may underestimate the accelerationrate of these competition camshafts. In thiscase you may find more accurate predic-tions from the Roller Lifter selection.

Camshaft Menu—Application-SpecificCamshafts

The second group within the Camshaftmenu contains five camshaft “grinds” thatare listed by application: 1) Stock Street/Economy Profile, 2) High PerformanceStreet Profile, 3) Dual Purpose Street/TrackProfile, 4) Drag-Race/Circle-Track Profile,and 5) Drag-Race High-Speed Profile. Anyof the three lifter types can be applied tothese cam profiles, adjusting the accelera-tion rates from mild to very aggressive.When any of these cam profiles are se-lected, the seat-to-seat IVO, IVC, EVO, andEVC are loaded into the Component Se-lection Box along with the camshaft de-scription (valve lift specs are calculated anddisplayed as described next). Simulationversions 2.5 and later also display the In-take Centerline, Intake Lobe Center Angle,Intake Duration, and Exhaust Duration forall simulated camshafts.

IMPORTANT NOTE: The intake andexhaust valve lifts for all application-spe-

cific cams are automatically calculated bythe simulation and displayed on screen fol-lowed by the term “(Auto)”. Valve lifts arebased on the valve-head diameters cho-sen by the user and displayed in the Cyl-inderhead category of the Component Se-lection Box. Note: If valve diameters arealso being automatically calculated—byselecting “Auto Calculate Valve Size” fromthe CylinderHead menu—a cylinder-borediameter and a cylinderhead selection mustbe made before the program can calculatethe valve diameters and, consequently, thevalve lifts. The simulation adjusts the in-take and exhaust valve lifts to maintainappropriate lift-to-diameter ratios for a widevariety of applications, from single-cylindersmall bore engines to large displacementracing engines with large valve diameters.The “auto calculation” feature for valve liftwill be suspended and, instead, permanentvalues used for any camshaft by choosingone of the two “Other” selections at thebottom of the Camshaft menu AFTERchoosing the desired camshaft (more onthis in the next section).

Stock Street/Economy Profile —Thisfirst profile is designed to simulate a typi-cal factory-stock cam. All cam timing eventsdisplayed in the Component Selection Boxare seat-to-seat measurements.

The EVO timing utilizes combustionpressure late into the power stroke andearly IVC minimizes intake flow reversion.Late IVO and early EVC produce only 22degrees of overlap, enough to harnesssome scavenging effects but restricted

Camshaft Type: Stock Street/Economy Lifter Type: Hydraulic

Cam Specs @: Seat To Seat Int Centerline: 115.0

IVO (BTDC): 12.0 IVC (ABDC): 62.0 Int Duration: 254.0

EVO (BBDC): 66.0 EVC (ATDC): 10.0 Exh Duration: 256.0

Int Lift @Valve: 0.XXX (Auto) Lobe-Center Angle: 116.5

Exh Lift @Valve: 0.XXX (Auto) Advance(+)/Retard(-): 0

This first menuprofile simulatesa typical factory-

stock cam. It isgenerally usedwith hydraulic

lifters.

This profilesimulates a high-

performancesolid-lifter camsimilar to ISKY

#201025.

Camshaft Type: High-Performance Street Lifter Type: Hydraulic







enough to prevent exhaust gas reversioninto the induction system. The characteris-tics of this cam are smooth idle, good powerfrom 1000 to 4500rpm, and good fueleconomy. This cam works well in high-torque demand applications. The StockStreet/Economy Profile cam is typicallyused with hydraulic lifters. As describedearlier, the intake and exhaust valve liftsfor all application-specific profiles are au-tomatically calculated by the simulation andare based on the valve diameters.

High Performance Street Profile —Thisprofile is designed to simulate a high-per-formance factory camshaft. All cam timingevents displayed in the Component Selec-tion Box are seat-to-seat measurements.

This camshaft uses relatively-late EVOto fully utilize combustion pressure andearly IVC minimizes intake flow reversion.IVO and EVC produce 62 degrees of over-lap, a profile that is clearly intended toharness exhaust scavenging effects. Themodestly-aggressive overlaps allow someexhaust gas reversion into the inductionsystem at lower engine speeds, affectingidle quality and low-speed torque. Thecharacteristics of this cam are fair idle, goodpower from 1500 to 6000rpm, and goodfuel economy. This cam develops consid-

erable power at higher engine speeds andis especially effective in lightweight ve-hicles. This High Performance Street Pro-file choice can be used with either hydrau-lic or solid lifters, and the simulation willaccurately model this cam with either lifterselection (choose hydraulic lifters for morestreet-oriented applications and solid liftersfor more high-performance oriented appli-cations). This cam is nearly identical to theISKY Hi-Rev Flat-Tappet cam part 201025for the smallblock Chevy. As describedearlier, the intake and exhaust valve liftsfor all application-specific profiles are au-tomatically calculated by the simulation andare based on the valve diameters.

Dual Purpose Street/Track Profile —This profile is designed to simulate a high-performance aftermarket camshaft. All camtiming events displayed in the ComponentSelection Box are seat-to-seat measure-ments.

EVO timing on this camshaft is begin-ning to move away from specs that wouldbe expected for simply utilizing combus-tion pressure with more of an emphasistoward early blowdown and minimizingexhaust pumping losses. The later IVCattempts to strike a balance between har-nessing the ram effects of the induction

The third menuchoice models adual-purposesolid-lifter camsimilar to ISKY#201281.

Camshaft Type: Dual-Purpose Street Lifter Type: Solid






Camshaft Type: Drag-Race High-Speed Lifter Type: Roller






Camshaft Type: Drag-Race Circle-Track Lifter Type: Solid






This profilesimulates acompetitioncamshaft similarto ISKY #201555.

This profilemodels an all-out cam. Theprofile is similarto ISKY Roller#201600.


system while minimizing intake flow rever-sion. IVO and EVC produce 64 degrees ofoverlap, a profile that is clearly designedto harness exhaust scavenging effects. Themodestly aggressive overlap can allowsome exhaust gas reversion into the in-duction system at lower engine speeds,affecting idle quality and low-speed torque.The characteristics of this cam are lopeyidle, good power from 2500 to 6500rpm,and modest fuel economy. This cam de-velops considerable power at higher en-gine speeds and is especially effective inlightweight vehicles. This Dual PurposeStreet/Track Profile choice can be used witheither hydraulic or solid lifters, and the simu-lation will accurately model this cam de-sign with either lifter selection (choosehydraulic lifters for more street-orientedapplications and solid lifters for more com-petition-oriented applications). The profileof this cam is close to the ISKY HydraulicSeries cam part 201281 for the smallblockChevy. As described earlier, the intake andexhaust valve lifts for all application-spe-cific profiles are automatically calculatedby the simulation and are based on thevalve diameters.

Drag-Race/Circle-Track Profile —Thisprofile is designed to simulate a competi-tion aftermarket camshaft. All cam timingevents displayed in the Component Selec-tion Box are seat-to-seat measurements.

EVO timing on this racing camshaftplaces less emphasis on utilizing combus-tion pressure and more emphasis on be-ginning early blowdown to minimize ex-haust pumping losses. The later IVC at-tempts to strike a balance between har-nessing the ram effects of the inductionsystem while minimizing intake flow rever-sion. IVO and EVC produce 90 degrees ofoverlap, a profile that is clearly intended tooptimize exhaust scavenging effects. Thisaggressive overlap is designed for openheaders and allows exhaust gas reversioninto the induction system at lower enginespeeds, affecting idle quality and torquebelow 3500rpm. The characteristics of thiscam are very lopey idle, good power from3600 to 7600rpm, with no consideration forfuel economy. This cam develops consid-

erable power at higher engine speeds andis especially effective in lightweight ve-hicles. This Drag-Race/Circle-Track Profilechoice can be used with either solid or rollerlifters, and the simulation will accuratelymodel this cam design with either lifterselection. The profile of this cam is similarto the ISKY Oval Track Flat Tappet Seriescam part 201555 for the smallblock Chevy.As described earlier, the intake and ex-haust valve lifts for all application-specificprofiles are automatically calculated by thesimulation and are based on the valve di-ameters.

Drag-Race High-Speed Profile —Thisprofile is designed to simulate an all-outcompetition aftermarket camshaft. All camtiming events displayed in the ComponentSelection Box are seat-to-seat measure-ments.

All timing events on this camshaft aredesigned to optimize power on large dis-

Several of the “generic” grinds that areincluded in the menu selection of the Fill-ing-And-Emptying simulation were se-lected from the ISKY CAMS catalog.ISKY’s catalog is “simulation-friendly,”listing seat-to-seat valve event timing fornearly every cam in their line.


placement engines at very high enginespeeds with large-tube, open headers, andhigh compression ratios. This camshaftmay not be effective in small displacementengines. EVO timing on this racing profileplaces the utilization of combustion pres-sure on the “back burner” and focusesemphasis on beginning early blowdown tominimize pumping losses during the ex-haust stroke. This technique will help powerat very high engine speeds, especially onlarge-displacement engines that do noteasily discharge the high volume of ex-haust gasses they produce. The late IVCattempts to harness the full ram effects ofthe induction system while relying on in-take pressure wave tuning to minimize in-take flow reversion. IVO and EVC produce104 degrees of overlap, a profile that isclearly intended to utilize exhaust scaveng-ing effects. This very aggressive overlapseriously affects idle quality and torque be-low 4000rpm. The characteristics of thiscam are extremely lopey idle, good powerfrom 4500 to 8500+rpm, with no consider-ation made for fuel consumption. This Drag-Race High-Speed Profile is designed to beused with roller lifters. The profile of thiscam is similar to the ISKY Roller Seriescam part 201600 for the smallblock Chevy.As described earlier, the intake and ex-haust valve lifts for all application-specificprofiles are automatically calculated by thesimulation and are based on the valve di-ameters.

Each of the above application-specificcams can be modified in any way by di-rectly entering valve-event or other cam-timing specs by choosing one of the two“Other” selections at the bottom of theCamshaft menu (more on this next).

Entering 0.050-Inch andSeat-To-Seat Timing

The last two selections in the Camshaftmenu are “Other” choices that allow thedirect entry of cam timing events to simu-late virtually any camshaft. The first choiceforces all specifications displayed and en-tered in the Camshaft category of the on-screen Component Selection Box to betreated as Seat-To-Seat timing specs. The

second choice instructs the simulation toassume that all cam timing values are0.050-inch timing specs. Whenever an“Other” selection is made that requires theprogram to switch from one timing methodto another, any currently displayed timingvalues are NOT changed; however, awarning message clearly indicates the tim-ing method has changed. In addition, thenew selected timing method is displayednext to “Cam Specs @:” in the Camshaftcategory. Furthermore, after an “Other” se-lection, the intake and exhaust valve-liftfields are switched to “(Man),” making en-tered or displayed valve-lift data perma-nent (turns off any auto-calculation tech-niques that may have been used to previ-ously calculate valve lift).

Only the four basic timing events areaffected by changes in seat-to-seat vs.0.050-inch measurement methods. As you’llnotice on the valve-motion curve drawing,changes in timing methods can only affectIVO, IVC, EVO, EVC, and the calculatedintake and exhaust duration. The remain-ing timing events, including Intake Center-line (ICA), Exhaust Centerline (ECA), LobeCenter Angle (LCA), and Intake and Ex-haust Valve Lift are not altered by eithermeasurement method because none ofthese specs are derived relative to any ofthe basic four valve events.

The seat-to-seat timing method mea-sures the valve timing—relative to pistonposition—when the valve or lifter has onlyjust begun to rise or has almost completelyreturned to the base circle on the closingramp. Unfortunately, there are no univer-sal seat-to-seat measuring standards.These are some of the more common:

0.004-inch valve rise for both intakeand exhaust0.006-inch valve rise for both intakeand exhaust0.007-inch open/0.010-close valve risefor both valves0.010-inch valve rise for both intakeand exhaust0.020-inch LIFTER rise for both intakeand exhaustThe timing specs measured using these

methods are meant to approximate theactual valve opening and closing points that


occur within the running engine. Becauseof this, seat-to-seat valve events are oftencalled the advertised or running timing. Aswe mentioned previously in this book, anengine simulation program needs just thistype of information to calculate the begin-ning and end of mass flow in the ports andcylinders, a crucial step in the process ofdetermining cylinder pressures and poweroutput. Because of this, seat-to-seat tim-ing specifications produce the most accu-rate simulation results. The 0.050-inch tim-ing figures, while accepted by Motion’ssimulation, must be internally converted toseat-to-seat figures, unfortunately a less-than-perfect process, before they can beused in the simulation.

In the early days of the “cam wars,”primarily during the 50’s and 60’s, the seat-to-seat timing method became popular withcam manufacturers as a way to “advertise”the duration of their popular grinds. Re-member the bigger-is-better axiom? It prob-ably reached its peak during this period. Atthat time, the way to be declared a winner

in the marketplace was to offer the “big-gest” and “baddest” camshaft, and thatmeant a cam with the longest duration.Manufacturers got so caught up in this fool-ishness that they used “trick” grindingmethods to extend the clearance rampsand artificially increase seat-to-seat dura-tion without appreciably affecting the valve-open duration (it was already too big). Bythis time, enthusiasts were confused by themyriad of seat-to-seat timing specs, andmany were installing camshafts incorrectly.Even worse, the very nature of seat-to-seattiming makes it difficult to “nail down” theprecise (rotational) position of the camduring engine assembly. To solve this prob-lem, cam manufacturers united (picturewater and oil!) to introduce a universal camspec primarily aimed at making cam instal-lation easier and more accurate. The newtechnique, called 0.050-inch timing, wasbased on the movement of the cam fol-lower (lifter) rather than the valve. Sincethe lifter is moving quite quickly at 0.050-inch it was easy to accurately index the

Seat-to-seat timingmeasures the valvetiming—relative to pistonposition—when the valveor (more rarely the lifter)has just begun to rise.Here dial indicators arepositioned on thevalvespring retainersand are measuring valverise, which is the mostcommon technique usedwith seat-to-seat timing(0.020-inch LIFTER riseis a notable exception).Timing specs measuredusing these methods aremeant to approximatethe actual valve openingand closing points thatoccur within the runningengine. Because of this,seat-to-seat valve eventsare often called theadvertised or runningtiming and will alwaysproduce the mostaccurate simulations.


cam to the crank position. Today, regard-less of who manufactures the cam, youwill always find 0.050-inch lifter rise timingpoints published on the cam card, simpli-fying cam installation.

The 0.050-inch lifter rise timing methodhas become one of the few standards inthe performance marketplace. In fact, somecam manufacturers have come to embracethis method so completely that they won’tpublish the old “advertised” or seat-to-seattiming events. This may go a long waytoward establishing a standard, but it’s areal step backwards from the standpoint oftesting cams in engine simulation programs.As we have said, an engine simulationprogram MUST know when the valves liftoff the seats and when they return to theirseats in order to calculate mass flow intoand out of the engine.

Let’s examine how much VALVE lifttypically occurs at 0.050-inch of LIFTERrise:

Valve Lift @ 0.050-inch Lifter Rise == (0.050 x Rocker Ratio) - Valve Lash

The 0.050-inch lifter risecam timing method hasbecome one of the fewstandards in the perfor-mance marketplace. Itmeasures the valvetiming—relative to pistonposition—when the lifterhas risen 0.050-inch offof the base circle of thecam. In the setuppictured here, the dialindicator is positionedon an intake lifter and isreading 0.050-inch; the0.050-inch valve timingpoint can now be readdirectly off of the degreewheel attached to thecrankshaft. Timing specs

measured using this method are not meant to approximate the actual valve openingand closing points, instead their purpose is to permit accurate cam installation. All0.050-inch timing specs entered into the Filling-And-Emptying program are internallyconverted to seat-to-seat timing. Because there is no way to precisely perform thisconversion, always try to obtain and use seat-to-seat event timing to optimize simula-tion accuracy.

For example (for 1.5 rockers and 30 thou-sandths lash):

Valve Lift @ 0.050 == 0.050 x 1.5 - 0.030= 0.045-inch

This example shows that the net valve liftis nearly equal to lifter rise. At this point,your knowledge of IC valve events shouldtell you that substantial flow occurs duringthis seemingly insignificant period. Considerthe EVO point for example. By 0.050-inch,the exhaust valve is well on its way todepressurizing the cylinder and havingdramatic effects on the power balance be-tween induced torque and pumping losseson the exhaust stroke. Furthermore, theEVO point blasts a pressure wave throughthe exhaust system that returns as a scav-enging wave during the overlap period.Similar critical functions occur at the othervalve timing points. Valve motion duringthe first 0.050-inch of lift cannot be disre-garded as insignificant when performingengine simulations. But what do you dowhen you just can’t find seat-to-seat timing


specs?The first answer to that question is pro-

vided in the second “Other” choice of theCamshaft pull-down menu (additional tipsfor finding “missing” valve event specs isprovided later in this guide). This selectiondeclares that all specifications displayedand entered in the program are assumedto have been obtained using the 0.050-inch Lifter-Rise timing method. (Wheneveran “Other” selection is made that requiresthe program to switch from one timingmethod to another, a warning messages isdisplayed indicating a change has beenmade in cam timing methods.) All 0.050-inch timing specs entered into the programare internally converted to seat-to-seat tim-ing points before the simulation is per-formed—remember, engine simulation issimply not possible without seat-to-seatvalve timing specs. Unfortunately, there isno precise way to make this conversion,so an estimation is performed of where theseat-to-seat points might lie based on theknown 0.050-inch timing points. Sometimesthe program is able to guess very closely,and the power curves will match dyno re-sults with that camshaft. At other times,the shape of the lobe is considerably dif-ferent than the program’s best guess, and

accuracy will suffer.We realize that this situation can be frus-

trating for some users, particularly whenyou consider that a few tech support peopleat cam manufacturers become belligerentwhen asked for seat-to-seat valve eventsor tell customers that the information is“proprietary.” Until these behind-the-timesindividuals “wake up” and realize that en-thusiasts have tools like engine simulationsand want to “test” cams before they buy,our best suggestion is to take your busi-ness elsewhere. But before you give upcompletely on your favorite cam grinder,take a look at the upcoming section oncalculating valve events. It may show youhow to calculate the needed timing fromthe jumble of specs printed in their camcatalog.

Camshaft Advance and Retard

In our earlier discussions on overlap,we found that this important cam timingevent can be adjusted in two ways: 1) thewrong way by changing the Lobe CenterAngle, or 2) the right way by changing theappropriate valve events to complementoverlap timing. Changing LCA alone effec-tively moves two valve events to improve

Installing offset cam bushing in the cam gear is a common method of advancing orretarding cam timing. While this method can improve power, it hurts almost as muchas it helps. Camshafts that show significant power gains from advanced or retardedtiming have the wrong event timing for the engine.


high-speed performance and the other twoevents decrease high-speed power. This“schizophrenic” tuning approach also oc-curs when the cam is advanced or retarded.

Selecting an “Other” choice from theCamshaft menu moves the cursor to theComponent Selection Box and allows thedirect entry of cam timing specifications.After you have entered all four valve eventsand both valve lift specs, the cursor movesto the “Advance(+)/Retard(-)” field. Chang-ing this spec from zero (the default) to apositive value advances the cam (in crankdegrees) while negative values retard thecam. The Advance/Retard function “shifts”all the intake and exhaust lobes the sameadvanced or retarded amount relative tothe crankshaft. Why would you want to dothis? The answer is simple: It is just aboutthe only “tuning” change available to theengine builder without regrinding or replac-ing the cam. While it’s possible to “tune”the cam using offset keys, special bush-ings, or multi-indexed sprockets, let’s in-

IVO LaterLow-Speed

EVC LaterHigh-Speed

EVO LaterLow-Speed

RetardedCam

EFFECTS OF RETARDING CAMSHAFT

IVC LaterHigh-Speed

ADVANCING/RETARDING A CAM IS ANOTHER“NONSENSE” TUNING APPROACH

.000

02040 4060100140180220260300 60 100 140 180 220 260 30020

.050

.100

.150

.200

.250

.300

.350

.400

.450

.500

INTAKE EXHAUST

BDC

TDC

BDC

The “schizophrenic” tuning approach we discussed earlier involving LCA also occurswhen the cam is advanced or retarded. If advancing or retarding the cam is actually“poor practice,” why is it so popular? The answer is simple: It is just about the only“tuning” change available to the engine builder without regrinding or replacing thecam. Advancing the cam slightly improves low-speed power, while retarding the camgives a small boost in high-speed power. If advancing or retarding allows the engineto perform better, the cam profile was not optimum in the first place.

vestigate what happens when all the valveevents are advanced or retarded from thecam manufacturer’s recommended timing.

It is generally accepted that advancingthe cam improves low-speed power whileretarding the cam improves high-speedpower. When the cam is advanced, IVCand EVC occur earlier and that tends toimprove low-speed performance; however,EVO and IVO also occur earlier, and thesechanges tend to improve power at higherengine speeds. The net result of theseconflicting changes is a slight boost in low-speed power. The same goes for retardingthe cam. Two events (later IVC and EVC)boost high-speed power and two (later EVOand IVO) boost low-speed performance.The net result is a slight boost in high-speed power.

Advancing or retarding a camshaft hasthe overall affect of reducing valve-timingefficiency in exchange for slight gains inlow- or high-speed power. Consequently,most cam grinders recommend avoiding


this tuning technique. If advancing or re-tarding allows the engine to perform betterin a specific rpm range, the cam profilewas probably not optimum in the first place.More power can be found at both ends ofthe rpm range by installing the right camrather than advancing or retarding thewrong cam. However, if you already own aspecific camshaft, slightly advanced orretarded timing may “fine tune” engineoutput to better suit your needs.

Calculating Valve Events AndUsing The Cam Math Calculator

Valve-motion curve drawings clearly in-dicate that cam timing events are related.No specific event is entirely isolated fromthe others. With sufficient information aboutthe derivative cam timing specs, it is pos-sible to calculate the basic four valve eventsneeded to run a simulation. What is theminimum amount of information needed tofigure out valve-event timing? There is noone answer to this question, but the an-swer that applies to the specs found inmany cam manufacturers’ catalogs is: In-take Duration, Exhaust Duration, LobeCenter Angle (LCA), and the Intake Cen-terline (ICA). With these four cam specs, itis possible to calculate the IVO, IVC, EVO,and EVC. The calculated valve opening andclosing events will be based on either seat-to-seat or 0.050-inch timing methods, de-pending on how the duration figures weremeasured. Remember, whenever you havea choice, always use seat-to-seat timingfigures; the simulation results will have thehighest accuracy. Also remember, the LCAand ICA do not change with timing meth-ods, so the same LCA and ICA values canbe used with either seat-to-seat or 0.050-inch duration figures.

Let’s use the following example andwork through the process of calculating thevalve events hidden within this cam timing.A particular cam catalog lists a smallblockFord cam as:Cam Type: HydraulicNet Valve Lift (Int): 0.448-inchNet Valve Lift (Exh): 0.472-inchDuration @ 0.050 (Int): 204-degreesDuration @ 0.050 (Exh): 214-degrees

Duration @ 0.006 (Int): 280-degreesDuration @ 0.006 (Exh): 290-degreesOverlap @ 0.006: 61-degreesIntake Centerline: 107-degreesLobe Center Angle: 112-degreesThis is a lot of information, but none of thebasic valve events are listed. Refer to thevalve-motion drawings. See if you can dis-cover the relationships described by thefollowing equations (they assume all camlobes are symmetric; while asymmetricprofiles are quite common, the amount ofasymmetry typically is very small, so thefollowing calculations should be accurateto within two or three degrees regardlessof the cam profile).The first step is to calculate how manydegrees before TDC the intake valve opens(when IVO occurs). This can be done withthe following formula (we’ll use seat-to-seatduration figures):Intake Valve Opening (IVO) =

= (Intake Duration / 2) - ICA= 280/2 - 107= 140 - 107= 33 degrees BTDCKnowing the IVO, it is possible to calcu-

late the intake valve closing point (IVC) bysimply subtracting the IVO from the intakeduration and then subtracting an additional180-degrees to account for the full intakestroke:Intake Valve Closing (IVC) =

= Intake Duration - IVO - 180= 280 - 33 - 180= 67 degrees ABDCWe’re halfway there. The next step re-

lates the known intake lobe timing to theexhaust lobe and then calculates the ex-haust timing events. We must first locatethe center of the exhaust lobe from theknown intake centerline using the lobecenter angle. The lobe center angle (LCA)is the distance (in cam degrees) betweenthe exact center of both lobes. Since weknow that the intake centerline (ICA) is thedistance from TDC to the center of theintake lobe, it is possible to calculate wherethe center of the exhaust lobe lies relativeto TDC:Exhaust Centerline (ECA) =

= (2 x Lobe Center Angle) - ICA= (2 x 112) - 107


= 224 - 107= 117 degreesNow that the exhaust centerline is

known, we can repeat the same processto find the exhaust valve-event timing:Exhaust Valve Closing (EVC) =

= (Exhaust Duration / 2) - ECL= (290/2) - 117= 145- 117= 28 degrees ATDCFinally, knowing the EVC, we can cal-

culate the Exhaust Valve Opening (EVO)point by subtracting the EVC from theexhaust duration and then subtracting an-other 180 degrees to account for the fullexhaust stroke:Exhaust Valve Opening (EVO) =

= Exhaust Duration - EVC - 180= 290 - 28 - 180= 82 deg. BBDCSo the valve events for this cam are:IVO = 33 degrees BTDCIVC = 67 degrees ABDCEVO = 82 degrees BBDCEVC = 28 degrees ATDC

If any of these valve events had turned outto be negative, which is possible for somestock-type cams measured using 0.050-inch timing, enter the calculated timing fig-ures into the simulation with the minus sign.For example, if EVC was determined to be

The series of calcula-tions performed onthese pages can bedone instantly by theCam Math Calculatorincluded in Motion’sFilling And Emptyingsimulation version 2.5(under development asthis book went to press;send in your registrationcard to find out moreabout upgrades and newproducts). By clickingon the MATH button inthe lower right of thescreen, the Cam Math

Calculator will “pop up” a window pre-loaded with the current cam timing. Any singleevent can be altered and the remaining events will be recalculated. This handyaddition to the program makes short work of entering cam data from somemanufacturer’s catalogs.

-10 degrees, this would mean that the valvecloses 10 degrees BEFORE TDC ratherthan after it. To tell the simulation that thisis the case, enter a -10 for EVC.

This series of calculations is performedautomatically by the Cam Math Calculatorincluded in Motion’s Filling And Emptyingsimulation version 2.5 (under developmentas this book went to press). By clicking onthe MATH button in the lower right of thescreen, the Cam Math Calculator will opena window and pre-load it with the cam tim-ing currently displayed on screen. If anysingle event is changed, the remainingevents are instantly recalculated and maybe saved to the main screen or discarded.This handy addition to the program makesshort work of not only entering cam datafrom manufacturer’s catalogs, but also youcan test the results of changes to LobeCenter Angle, Intake Centerline, IntakeDuration and Exhaust Duration. Combinedwith the ability to change IVO, IVC, EVO,EVC, and overall advance and retard fromthe main screen, the new version 2.5 withthe Cam Math Calculator allows changingvirtually EVERY cam timing event andmeasuring its result.


Appendix-A Common Questions

COMMONLY ASKED QUESTIONS

The following information may be help-ful in answering questions and solving prob-lems that you encounter installing and usingthe DeskTop Dyno. If you don’t find ananswer to your problem here, send in theMail/Fax Tech Support Form provided onpage 97 and in the Installation AndQuickStart guide provided with your soft-ware. We will review your problem andreply to you as soon as possible.

Question: Received an “Error ReadingDrive A (or B)” message whenattempting to run or install theDeskTop Dyno. What does thismean?

Answer: This means DOS cannot readthe disk in your floppy drive.The disk may not be fullyseated in your drive or the drivedoor (or lock arm) may not befully latched. If you can prop-erly read other disks in yourdrive, but the DeskTop Dynodistribution disk produces errormessages, try requesting a di-rectory of a known-good diskby entering DIR A: or CHKDSKA: and then perform thosesame operations with the Desk-Top Dyno disk. If these opera-tions produce an error mes-sage only when using theDeskTop Dyno disk, the diskis almost certainly defective.Return the disk to Motion Soft-ware, Inc., for a free replace-ment (address at bottom ofTech Support Form).

Question: Encountered a “DeskTop DynoHas Not Been Properly In-stalled...” error message whentrying to run the DeskTopDyno. Why?

Answer : This means that files requiredby the DeskTop Dyno were notfound on your system. Reinstall

the DeskTop Dyno using theSETUP program from the origi-nal distribution disk.

Question: The results of the simulationare listed in a simple chart onscreen. Why are no powercurves displayed?

Answer : Any monitor and display cardwill work with the DeskTopDyno, however, systems withEGA or high-resolution graph-ics capability will allow the dis-play of horsepower and torquecurve graphics. If you do nothave an EGA or better graph-ics system (you are using aCGA or another low-resolutiondisplay), an on-screen “chart”listing for horsepower andtorque will be substituted forpower curves.

Question: What are the atmospheric andenvironmental conditions as-sumed by the program to pre-dict horsepower?

Answer: Motion’s Filling-And-Emptyingsoftware closely simulates theconditions that exist during anactual engine dyno test. Thegoal is to reliably predict thetorque and horsepower that adynamometer will measurethroughout the rpm range whilethe engine and dyno are run-ning through a programmedtest. However, engine powercan vary considerably from onedyno test to another if environ-mental and other critical condi-tions are not carefully con-trolled. In fact, many of thediscrepancies between dynotests are due to variabilities inwhat should have been “fixed”conditions. Among the many in-terviews conducted during theresearch and development ofthe software and for this book,dyno operators and engine


owners readily acknowledgedthe possibilities of errors inhorsepower measurements;however, very few these indi-viduals believe that they hadbeen lead astray by erroneousdyno readings. It seems thatwhen test numbers “pop up” ona computer screen after aground-pounding dyno test, thehorsepower and torque valueswere accepted as “gospel.” Inother words, “dynos do lie, butthey wouldn’t do that to me”!Unless the dyno operator andtest personnel are extremelycareful to monitor and controlthe surrounding conditions, in-cluding calibration of the instru-mentation, dyno measurementsare nearly worthless. Control-ling these same variables in anengine simulation program isinfinitely easier, but neverthe-less just as essential. Initialconditions of temperature,pressure, energy, and method-ology must be established andcarefully followed. Here aresome of the assumptions withinthe simulation software thatestablish a modeling baseline:Fuel:1) Gasoline rated at 19,000btu/lb as a standard fuel2) The fuel is assumed to havesufficient octane to preventdetonation.3) The air/fuel ratio is alwaysmaintained at 13:1 for optimumpower.Environment:1) Air for induction is 68-de-grees (F), dry (0% humidity),and of 29.86-in/Hg atmosphericpressure.2) The engine, oil, and cool-ant have been warmed to op-erating temperature.

Methodology:1) The engine is put through aseries of “step” tests. That isthe load is adjusted to “holdback” engine speed as thethrottle is opened wide. Thenthe load is adjusted to allowthe engine speed to rise to thefirst test point, 2000rpm in thecase of the simulation. Theengine is held at this speed fora few seconds and a powerreading is taken. Then enginespeed is allowed to increaseto the next step, 2500rpm, anda second power reading istaken. This process is contin-ued until the maximum testingspeed is reached.2) Since the testing proceduretakes the engine up in 500rpmsteps, and it is held steadyduring the measurement, themeasured power does not re-flect any losses from acceler-ating the rotating assembly(crank, rods, etc.) as would befound in a “real-world” applica-tion, such as drag racing.

Question: When I choose a carburetorthat is too large for an engine(for example 1200cfm on a 283Chevy), why does the powerincrease without the typicallyseen “bog” at low speeds?

Answer : The DeskTop Dyno, along withvirtually any current computersimulation program, cannotmodel over-carburetion andshow the usual reduction inlow-end performance that thiscauses. In reality, carburetorsthat are too large for an en-gine develop fuel atomizationand air/fuel ratio instabilities.Unfortunately, this phenom-enon is extremely complicatedto model. The DeskTop Dynoassumes an optimum air/fuel



ratio regardless of the selectedCFM rating. While you can getpositive results from larger-and-larger induction flows (by theway, this is not far from realitywhen optimum air/fuel ratioscan be maintained, as is thecase in electronic fuel-injectionsystems), you can't go wrongif you use common sense whenselecting induction/carburetorflow capacities.

Question: I built a relatively stock enginebut installed a drag-race cam-shaft. The engine only pro-duced 9 hp @ 2000 rpm. Isthis correct?

Answer : Yes. Very low power outputsat low engine speeds occurwhen radical camshafts areused without complementarycomponents, such as high-flowcylinderheads, high compres-sion ratios, and exhaust sys-tem components that match theperformance potential of thecam. In fact, some low perfor-mance engines with radicalcamshafts will show zerohorsepower at low speeds. Thismeans that if the engine wasassembled and installed on adynamometer, it would not pro-duce enough power to offsetany measurable load.

Question: The power predicted with theDeskTop Dyno seems too highfor the Oldsmobile 455 enginethat I am testing. Why?

Answer : You selected cylinderheads foryour Olds engine from the BigBlock choices in theCylinderHead menu. While thismay seem like the logicalchoice, some big-block enginesare equipped with heads thatflow significantly less air thanother large (big-block) engines.You will obtain more accurate

power predictions by choosingone of the Smallblock headselections when building one ofthese large-displacement en-gines. In addition toOldsmobile, versions of the429-460 Ford engines alsocame with restrictive cylinder-heads. When you make cylin-derhead choices, keep thesetips in mind: Restrictive ap-plies to small-port heads (e.g.,low-performance engines, suchas early 260cid Ford enginesor even flatheads) where theports are small and restrictivefor the installed valve sizes;consider Smallblock to de-scribe heads with ports that aresized adequately or with per-formance in mind; Bigblockapplies to heads where theports are large or the valvesare canted for improved flow(the first three big-block choicesmodel oval-port heads, the lasttwo choices simulate rectangu-lar-port heads); and 4-ValveHeads apply to many importand some racing configura-tions.

Question: The DeskTop Dyno producedan error message “THE SE-LECTED COMBINATIONPRODUCED A CALCULATIONERROR...” What went wrong?

Answer : The combination of compo-nents you have selected pro-duced a calculation error in thesimulation process (this is usu-ally caused by a mathematicalinstability as the program re-peatedly performed calcula-tions to “home in” on a vari-able—called iteration). This isoften caused by using restric-tive induction flow on large-dis-placement engines, or by us-ing radical cam timing on oth-


erwise mild engines. Try reduc-ing the EVO timing specs, in-creasing the induction flow,selecting a cam with less du-ration, or reducing the com-pression ratio. A balancedgroup of components shouldnot produce this error.

Question: The DeskTop Dyno takes 15minutes to complete a simula-tion and draw horsepower andtorque curves. Is there a prob-lem with my computer or thesoftware?

Answer : Your computer does not havea math coprocessor (speeds upthe simulation from 50 to 300times). The DeskTop Dynouses a powerful full-cycle simu-lation that performs millions ofcalculations for each point onthe power curves, and thistakes some time. But what youlose in speed with the Desk-Top Dyno, you gain in accu-racy over less sophisticatedprograms.

Question: How can I stop the simulationcalculation if I realize that I'vemade a mistake selecting acomponent so that I don't haveto wait for the program to draw

the complete power and torquecurves?

Answer: You can halt a simulation bypressing the ESCape key.Pressing Enter will resume thecalculation; however, a secondpress of the ESCape key willabort the simulation run and re-turn to the Main ProgramScreen.

Question: The DeskTop Dyno calculatedthe total Combustion Volumeat 92ccs. But I know my cylin-derheads have only 75ccs.What's wrong with the Desk-Top Dyno?

Answer : Nothing. The confusion comesfrom assuming that the calcu-lated Total Combustion Volumeis the same as your measuredcombustion-chamber volume.The Total Combustion Volumeis the entire volume that re-mains when the piston reachestop dead center. This includesthe combustion chamber, theremaining space above the pis-ton top and below the deck sur-face, and the valve pockets; butit excludes any portion of thepiston that protrudes into thecombustion chamber. Com-

When severely unbal-anced components areselected from the menusor entered manually, theiterative process ofsimulation can fall intomathematical instabili-ties, typically causing thecurves to form a ragged“sawtooth” plot. Simplyselecting a more bal-anced combination ofparts and componentspecs will eliminateinstabilities.



pression ratio is calculated bythis formula:

Cyl/Volume + Comb/VolCR = —————————————

Comb/Volume

Both cylinder volume and com-bustion volume are calculatedand displayed in the Compo-nent Selection Boxes.

Question: The horsepower producedwhen I enter the seat-to-seattiming on my cam card doesnot match the horsepowerwhen I enter the 0.050-inch tim-ing figures for the same cam-shaft. Why are there differ-ences?

Answer : The DeskTop Dyno uses thetiming specs found on yourcam card, and in cammanufacturer's catalogs, to de-velop a valve-motion curve(and from this, develops theinstantaneous airflow for eachport). Neither the seat-to-seatnor 0.050-inch timing figuresprecisely describe actual valvemotion; you would need tomeasure valve position at eachdegree of crank rotation tocome close to developing anexact valve-motion diagram!Lacking this, the DeskTopDyno “creates” its own valve-motion diagram for use in latercalculations of power andtorque. A lot can happen in in-duction airflow between thetime the valve is on the seatand when is reaches 0.050-inch of lifter rise. It is impos-sible (without exact measure-ment) to precisely “guess” theshape of the cam or the mo-tion of the valve. This is thereason for the calculated dif-ferences. When in doubt, use

seat-to-seat timing figures.They provide the DeskTopDyno more information aboutvalve motion, and are morelikely to produce accurate simu-lated power levels.

Question: When I build an engine, themenus close after each com-ponent selection. Why don’t themenus open, one-after-the-other, until the entire engine isassembled?

Answer : This would make engine as-sembly from scratch easier. Butit gets in the way when youwant to change individual com-ponents to evaluate thechanges in torque and horse-power. And since the truepower of the DeskTop Dynolies in its ability to make back-to-back tests—and it’s theseback-to-back tests that are es-sential steps in finding the bestcombination—we designed themenus to optimize changing in-dividual components andquickly running back-to-backtests.

Question: How does the DeskTop Dynoallow for hydraulic, solid, androller lifters?

Answer : The DeskTop Dyno calculatesa valve-motion diagram that isused in the subsequent calcu-lations to predict horsepowerand torque. When the choiceis made to move from hydrau-lic to solid, and then from solidto roller lifters, the DeskTopDyno increases the valve ac-celeration rates to coincide withthe lobe shapes that are com-monly found on these camgrinds. It is impossible (with-out exact measurement ofvalve position at each degreeof crank motion) to “guess” theprecise shape of the cam or


the motion of the valve, but theDeskTop Dyno has proven tobe remarkably accurate insimulating real-world valvemotion and horsepower.

Question: Can I change rockerarm ratioswith the DeskTop Dyno?

Answer : Yes. Simply use this formulato alter valve lift (the DeskTopDyno will calculate the newvalve motion throughout the liftcurve):

New RatioNew Lift = Old Lift x ——————

Old Ratio

When you have calculated thenew maximum valve lifts for theintake and exhaust valves,enter these numbers directlyinto the Component SelectionsBox by choosing one of theOTHER choices in the Cam-shaft menu.

Question: Why can’t I just enter my portflow? Wouldn’t that result inmore accurate power predic-tions?

Answer : If you could input the flow ofboth the intake and exhaustports at valve lifts from zero tomaximum lift for each 0.010-inch of valve motion, yes. Pub-lished flow figures for ports arevirtually meaningless when itcomes to predicting power po-tential. This is because thepeak port flow only occurswhen the valve reaches somemaximum lift figure specified bythe head manufacturer (andmaximum lift in your enginemay not be the same as themaximum lift that was used torate port flow). Furthermore,the valve is only at maximumlift for a small fraction of its mo-tion. To predict power you need

to know the flow at each pointof valve lift. This is calculatedby the DeskTop Dyno from: 1)the valve head diameter, 2)knowing how the port shape(restrictive, smallblock, big-block, etc.) will affect flowacross the valve, and 3) thevalve motion diagram calcu-lated from the valve-event tim-ing.

Question: I found the published factoryseat-to-seat valve timing forPontiac engine that I am build-ing, but I can't enter the valveevents into the DeskTop Dyno.The IVC occurs at 110 degrees(ABDC), and I can only enterup to 100 in the program.

Answer : There are so many ways thatcam specs can be describedfor cataloging purposes that it'sconfusing to anyone trying toenter timing specs into an en-gine simulation program. YourPontiac is a classic example ofthis lack of standards. ThePontiac cam listed in the fac-tory manual is a hydraulic grindwith seat-to-seat timing mea-sured at 0.001-inch valve rise.Because the cam is designedfor long life and quite opera-tion, it has shallow openingramps. This is the reason forthe large number of crank de-grees between the opening andclosing points. In fact, duringthe first 35 degrees of crankrotation, the lifter rises less than0.010-inch. If this wasn't thecase and the valve opened andclosed at the specified timingpoints listed in the factorymanual, the cam would haveover 350-degrees duration, andit's unlikely the engine wouldeven start! The DeskTop Dynocan use 0.004- or 0.006-inch



valve rise, 0.007-open/0.010-close valve rise, or even 0.020-inch lifter rise for seat-to-seattiming. But the 0.001-inch fig-ures published in your factorymanual are useless for enginesimulation purposes.

Question: My cam manufacturer's cata-log does not list seat-to-seatvalve-event timing. But it doeslist seat-to-seat intake and ex-haust duration, lobe-centerangle, and intake centerline.Can I calculate the valve-eventtiming from these figures?

Answer : Yes. To calculate the intakeand exhaust opening and clos-ing points, you must have allof the following information:1) Intake Duration2) Exhaust Duration3) Lobe-Center Angle (some-times called lobe separationangle).4) And the Intake CenterlineAngle .To perform the required calcu-lations refer to the step-by-stepprocedure described on pages86 and 87. Note: Version2.5.7M of the DeskTop Dyno(under development when this

guide was published) incorpo-rates a Cam Math Calculatorthat performs this calculationinstantly.

Question: I have been attempting to testcamshafts from a listing in acatalog. I can find the durationand lobe center angle. The cammanufacturer won't give me theseat-to-seat timing (they act likeit's a trade secret). Can I usethe available data to test theircams?

Answer : No. As stated in the previousanswer, you also need the in-take-center angle to relate camlobe positions to TDC and,therefore, crank position. Freelyproviding seat-to-seat timing orany of the other cam specsdiscussed in this booklet posesno threat to any cam grinder. Ittakes a lot more than valve-event timing to manufacture aquality cam; full profiles of thelobes are needed to ensuremechanically and dynamicallystable operation. Cam compa-nies that refuse to provide po-tential customers with simplevalve-event information forevaluation in programs like the

The Cam Math Calculatoris included in Motion’sFilling And Emptyingsimulation version 2.5. Byclicking on the MATHbutton the Cam MathCalculator will “pop up” awindow pre-loaded withthe current cam timing.Any single event can bealtered and the remainingevents will be recalcu-lated.


DeskTop Dyno are simply liv-ing in the “dark ages.” Our sug-gestion is to contact anothercam manufacturer (look intoISKY cams; see page 80).

Question: Everyone talks about howlonger rods make more power.Why isn’t rod length one of thechoices in the pull-downmenus?

Answer : Tests we have performed withthe DeskTop Dyno show thatrod length has virtually no af-fect on power. We realize thatmany actual dyno tests haveshown power increases, butour simulation tests tell us thatthe power, when found, prob-ably has little to do with pistondwell at TDC (and the associ-ated thermodynamic effects) orchanges in rod angularity onthe crank pin. The measuredpower differences are mostlikely due to a reduction of fric-tion on the cylinderwall fromchanges in side-loading on thepiston. This can vary with borefinish, ring stability, pistonshape, the frictional propertiesof the lubricant, etc. These vari-abilities are highly unpredict-able. Some development, afterall, can only be done in the realworld on a engine dynamom-eter.

Question: Can I print out the horsepowerand torque curves? I've triedusing PrintScreen and all I getare the number, no curves.

Answer: You can print a dyno test sheetby selecting “P” from the Simu-lation Completed Choicesprompt after any simulation run.The DeskTop Dyno printoutincludes a listing (in chart form)of the exact horsepower andtorque figures and a completesummary of the components

selected for that test. Becauseof the limitations in DOS (theDeskTop Dyno operates underDOS even while running inWindows) graphics printing ofthe power curves is not directlysupported. However DOS andWindows users can still obtaina printout of the SimulatorScreen including the horse-power and torque curves. Hereare some tips and hints thatwill help you obtain printedgraphic output.Printing under DOS or Win-dows 3.xYou can print the SimulatorScreen, including powercurves, on almost any IBM-compatible, graphics-capableprinter using the DOS utilitycalled GRAPHICS.COM. If youare using DOS version 5.0 orlater, you will find this nifty pro-gram in your DOS subdirectory.Type GRAPHICS at the DOSprompt before you start theDeskTop Dyno to enhance thePrint-Screen function to includegraphics printing. Then, withthe curves displayed on screenthat you would like to print, si-multaneously press the Shiftand PrintScreen keys to sendthe current graphics screen toyour printer.GRAPHICS has several op-tions that “fine tune” the com-mand for various printers andoutput styles. Here are a fewof the more useful methods ofusing GRAPHICS (refer to yourDOS manual for a complete listof options):Type at the DOS prompt:GRAPHICS LASERJETorGRAPHICS LASERJETIIto tell the GRAPHICS com-



mand that the output will beprinted on a HP or compatiblelaser printer.Type at the DOS prompt:GRAPHICS LASERJET /Rto print the image as it appearson screen rather than reversed.Reversed (black to white) print-ing is the default.Printing under Windows 95Windows 95 incorporates DOSversion 7.0 and is not supplied(nor is compatible) with theGRAPHICS command. Instead,when a simulation is com-pleted, press the ALT andENTER keys together to switchthe Simulator Screen from full-screen to a window. Then se-lect and copy the screen, in-cluding the power curve graph,to the clipboard (see the Win-dows help system for instruc-tions on how to accomplishthis). Open the Windows Paintprogram and paste the clip-board into Paint. Then, simplyprint the image to your Win-dows printer.

Question: I have an NEC Ready com-puter system. When I run asimulation, the curves are dis-played properly, but the num-bers and other words surround-ing the power curves is garbledand completely unreadable.How can I fix this?

Answer : The NEC Ready series of com-puters was released for sale inthe U.S. during the 1995 Christ-mas sale season. This systemhas a “bug” in firmware (un-changeable, internal codestored in ROM chips) thatcauses it to improperly displayEGA graphics characters. Itmaps the lower 128 ASCII totheir upper 128 counterparts.Motion Software has developed

a fix for this problem (a VGAversion of the power curvesscreen) that is available at nocharge. Return your DeskTopDyno disk, along with a notementioning your NEC computerproblem, to: Motion Software,Inc., NEC replacement, 535West Lambert, Bldg. E, Brea,CA, 92821.

Question: I have tried many different en-gine combinations using thesame engine displacementsand have noticed that severalof the power curves begin atnearly the same horsepowerand torque values at 2000rpm.Why are they so similar at thisengine speed?

Answer : Since the DeskTop Dyno usesa simulation technique that it-erates toward an answer—per-forms a series of calculationsthat approach a more and moreaccurate result—the first powerpoint must be developed basedon educated “guesses” aboutmass flow and other variables.The next point, at 2500rpm, iscalculated from the startingpoint, plus the data obtainedfrom the completed simulation,so accuracy is higher. By3000rpm, the power points arebased on simulation calcula-tions with virtually no remain-ing influence from the initial es-timations.


Please use this form (or a copy) to obtain technical support for the DeskTop Dyno fromMotion Software, Inc. Fill out all applicable information about your system configuration,program version, and describe your problem as completely as possible. We will attemptto duplicate the problem and respond to your question. Mail or fax this form to theaddress below.

Mail/Fax Tech Support Form For DeskTop Dynoop Dyno

(DeskTop Dynos Mini Guide)

Your Phone ( ) ______-____________ Your Fax ( ) ______-_____________

Your Name __________________________________________________________________

Address _________________________________________ Apt. or Building ___________

City _____________________________________ State ______ ZipCode _____________

Brand of computer _____________________________ CPU __________ Speed _______

Floppy ❏—5-1/4 ❏—3-1/2 Size of hard drive ______________ Amt of RAM ________

Describe your monitor and graphics card ______________________________________

____________________________________________________________________________

Version of DOS (enter the command VER at the DOS prompt) _____________________

Version of DeskTop Dyno (refer to distribution disk label) _________________________

List the TSR (Terminate and Stay Resident) programs loaded by your Autoexec.bat orConfig.sys files that are present when you are running the DeskTop Dyno.

_____________________________________________________________________________

_____________________________________________________________________________

_____________________________________________________________________________

_____________________________________________________________________________

Please describe the problem, the point in the program at which the problem firstoccurred and, if necessary, the menu choices that caused the problem to occur.

_____________________________________________________________________________

_____________________________________________________________________________

_____________________________________________________________________________

_____________________________________________________________________________

_____________________________________________________________________________

Can you duplicate the problem? _______________________________________________

Mail this form to: Motion Software, Inc., 535 West Lambert, Bldg. E, Brea, CA 92821

Or fax to: 714-255-7956 (24 hours)


Appendix-B Glossary

0.050-Inch Cam Timing Method —See CamTiming, @ 0.050-inch .ABDC or After Bottom Dead Center —Anyposition of the piston in the cylinder bore af-ter its lowest point in the stroke (BDC). ABDCis measured in degrees of crankshaft rota-tion after BDC. For example, the point atwhich the intake valve closes (IVC) may beindicated as 60-degrees ABDC. In otherwords, the intake valve would close 60 de-grees after the beginning of the compressionstroke (the compression stroke begins atBDC).Air-Fuel Ratio —The proportion of air to fuel—by weight—that is produced by the carbure-tor or injector.ATDC or After Top Dead Center —Any po-sition of the piston in the cylinder bore afterits highest point in the stroke (TDC). ATDC ismeasured in degrees of crankshaft rotationafter TDC. For example, the point at whichthe exhaust valve closes (EVC) may be indi-cated as 30-degrees ATDC. In other words,the exhaust valve would close 30 degreesafter the beginning of the intake stroke (theintake stroke begins at TDC).Atmospheric Pressure —The pressure cre-ated by the weight of the gases in the atmo-sphere. Measured at sea level this pressureis about 14.69psi.Back Pressure —A pressure developed whena moving liquid or gaseous mass passesthrough a restriction. “Backpressure” oftenrefers to the pressure generated within theexhaust system from internal restrictions fromtubing and tubing bends, mufflers, catalyticconverters, tailpipes, or even turbochargers.BBDC or Before Bottom Dead Center —Anyposition of the piston in the cylinder bore be-fore its lowest point in the stroke (BDC).BBDC is measured in degrees of crankshaftrotation before BDC. For example, the pointat which the exhaust valve opens (EVO) maybe indicated as 60-degrees BBDC. In otherwords, the exhaust valve would open 60 de-grees before the exhaust stroke begins (theexhaust stroke begins at BDC).Big-Block —A generic term that usually re-fers to a V8 engine with a displacement that

is large enough to require a physically “big-ger” engine block. Typical big-block enginesdisplace over 400 cubic inches.Blowdown or Cylinder Blowdown —Blow-down occurs during the period between ex-haust valve opening and BDC. It is the pe-riod (measured in crank degrees) duringwhich residual exhaust gases are expelledfrom the engine before the exhaust strokebegins. Residual gasses not discharged dur-ing blowdown must be physically “pumped”out of the cylinder during the exhaust stroke,lowering power output from consumed “pump-ing work.”Bore or Cylinder Bore —The internal surfaceof a cylindrical volume used to retain andseal a moving piston and ring assembly.“Bore” is commonly used to refer to the cylin-der bore diameter, unusually measured ininches or millimeters. Bore surfaces are ma-chined or ground precisely to afford an opti-mum ring seal and minimum friction with themoving piston and rings.Brake Horsepower (bhp) —Brake horse-power (sometimes referred to as shaft horse-power) is always measured at the flywheel orcrankshaft by a “brake” or absorbing unit.Gross brake horsepower describes the poweroutput of an engine in stripped-down, “race-ready” trim. Net brake horsepower measuresthe power at the flywheel when the engine istested with all standard accessories attachedand functioning. Also see Horsepower, Indi-cated Horsepower, Friction Horsepower, andTorque.Brake Mean Effective Pressure (bmep) —Atheoretical average pressure that would haveto be present in each cylinder during thepower stroke to reproduce the force on thecrankshaft measured by the absorber (brake)on a dynamometer. The bmep present dur-ing the power stroke would produce the samepower generated by the varying pressures inthe cylinder throughout the entire four-cycleprocess.BTDC or Before Top Dead Center —Anyposition of the piston in the cylinder bore be-fore its highest point in the stroke (TDC).BTDC is measured in degrees of crankshaft


rotation before TDC. For example, the pointat which the intake valve opens (IVO) maybe indicated as 30-degrees BTDC. In otherwords, the intake valve would open 30 de-grees before the intake stroke begins (theintake stroke begins at TDC).Cam Timing, @ 0.050-Lift —This method ofdetermining camshaft valve timing is basedon 0.050 inches of tappet rise to pinpoint tim-ing events. The 0.050 inch method was de-veloped to help engine builders accuratelyinstall camshafts. Lifter rise is quite rapid at0.050-inch lift, allowing the cam to be pre-cisely indexed to the crankshaft. Camshafttiming events are always measured in crank-shaft degrees, relative to TDC or BDC.Cam Timing, @ Seat-To-Seat —This methodof determining camshaft timing uses a spe-cific valve lift (determined by the cam manu-facturer) to define the beginning or ending ofvalve events. There is no universally acceptedvalve lift used to define seat-to-seat cam tim-ing, however, the Society of Automotive En-gineers (S.A.E) has accepted 0.006-inch valvelift as its standard definition. Camshaft timingevents are always measured in crankshaftdegrees, relative to TDC.Camshaft Advance/Retard — This refers tothe amount of advance or retard that the camis installed from the manufacturers recom-mended setting. Focusing on intake timing,advancing the cam closes the intake valveearlier. This setting typically increases low-end performance. The retarded cam closesthe intake valve later which tends to help topend performance.Camshaft Follower or Lifter —Usually ametal cylinder (closed at one end) that rubsagainst the cam lobe and converts the rotarymotion of the cam to an up/down motion re-quired to open and close valves, operate fuelpumps, etc. Cam followers (lifters) can incor-porate rollers, a design that can improve reli-ability and performance in many applications.Roller lifters are used extensively in racingwhere valve lift and valve-lift rates are veryhigh, since they can withstand higher dynamicloads. In overhead cam engines, the cam fol-lower is usually incorporated into the rocker

arm that directly actuates the valve; in thisdesign push rods are eliminated.Camshaft Grind —The shape of the camlobe. Determines when the intake and ex-haust valves open and close and how highthey lift off of the seats. The shape also de-termines how fast the valves open and close,i.e., how much acceleration the valves andsprings experience. High acceleration ratecams require large-diameter solid, mushroom,or roller lifters.Camshaft Lift— The maximum height of thecam lobe above the base-circle diameter. Ahigher lobe opens the valves further, oftenimproving engine performance. Lobe lift mustbe multiplied by the rocker ratio (for enginesusing rocker arms) to obtain total valve lift.Lifting the valve more than 1/3 the head di-ameter generally yields little additional per-formance. Faster valve opening rates addstress and increase valvetrain wear but canfurther improve performance. High lift ratesusually require specially designed, high-strength components.Camshaft Lobe —The eccentrically shapedportion of a camshaft on which the camshaftfollower or lifter rides. The shape of intakeand exhaust cam lobes are important enginedesign criterion. They directly affect engineefficiency, power output, the rate (how fast)the valves open and close, and control valve-train life and maximum valvetrain/engine rpm.Camshaft Timing —The rotational position ofthe camshaft, relative to the crankshaft, i.e.,the point at which the cam lobes open andclose the valves relative to piston position.Two common methods are used to indicatethe location of valve events: the Seat-To-Seatand 0.050-inch timing methods. For simula-tion purposes, Seat-To-Seat timing valuesyield more accurate horsepower and torquepredictions. Camshaft timing can be adjustedby using offset keys or offset bushings (or byredesigning the cam profile). Valve-to-pistonclearance will vary as cam timing is altered;always ensure that adequate clearance ex-ists after varying cam timing frommanufacturer’s specifications. See Cam Tim-ing @ Seat-To-Seat and Cam Timing @


Appendix-B Glossary

0.050-Inch.Carburetor —A device that combines fuel withair entering the engine; capable of precisioncontrol over the air volume and the ratio ofthe fuel-to-air mixture.Centerline —An imaginary line runningthrough the center of a part along its axis,e.g., the centerline of a crankshaft runningfrom front-to-back directly through the centerof the main-bearing journals.Closed Headers or Closed Exhaust Sys-tem—Refers to an exhaust system that in-cludes mufflers; not open to the atmosphere.Combustion Chamber or CombustionChamber Volume —The volume containedwithin the cavity or space enclosed by thecylinderhead, including the “top” surfaces ofthe intake and exhaust valves and the sparkplug. Not the same volume as the combus-tion space volume.Combustion Space or Combustion SpaceVolume —The volume contained within thecylinderhead, plus (or minus) the piston dome(or dish) volume, plus any volume displacedby the compressed head gasket, plus (or mi-nus) any additional volume created by thepiston not fully rising to the top of the bore(or extending beyond the top of the bore) ofthe cylinder at TDC. This volume is used tocalculate compression ratio.Compression Pressure —The pressure cre-ated in the cylinder when the piston movestoward top dead center (TDC) after the in-take valve closes, trapping the induced charge(normally a fuel/air mixture) within the cylin-der. Compression pressure can be measuredby installing a pressure gauge in the cylinderin place of the spark plug and “cranking” theengine with the starter motor. To improvemeasurement accuracy, the throttle is usu-ally held wide open and the remaining sparkplugs are removed to minimize cranking loadsand optimize pressures in the cylinder undertest.Compression Ratio —The ratio of the totalvolume enclosed in a cylinder when the pis-ton is located at BDC compared to the vol-ume enclosed when the piston is at TDC (vol-ume at TDC is called the combustion space

volume). The formula to calculate compres-sion ratio is: (Swept Cylinder Volume + Com-bustion Space Volume)/Combustion SpaceVolume = Compression Ratio.Compression Stroke —One of the four 180-degree full “sweeps” of the piston moving inthe cylinder of a four-stroke, internal-combus-tion engine (originally devised by NikolausOtto in 1876). During the compression stroke,the piston moves from BDC to TDC and com-presses the air/fuel mixture. Note: The 180-degree duration of the compression stroke iscommonly longer than the duration betweenthe intake valve-closing point and top deadcenter (or ignition), sometimes referred to asthe true “Compression Cycle.” The compres-sion stroke is followed by the power stroke.Cubic Inch Displacement or CID—Theswept volume of all the pistons in the cylin-ders in an engine expressed in cubic inches.The cylinder displacement is calculated withthis formula: (Bore x Bore x Pi x Stroke xNo.Cyl.)/4. When the bore and stroke aremeasured in inches, the engine displacementcalculated in cubic inches.Cylinder and Cylinder Bore —The cylinderserves three important functions in an inter-nal-combustion (IC) engine: 1) retains the pis-ton and rings, and for this job must be pre-cisely round and have a uniform diameter(for performance applications 0.0005-inch tol-erance is considered the maximum allowable);2) must have a surface finish that ensuresboth optimum ring seal (smooth and true) andyet provides adequate lubrication retention toensure long life for both the piston and rings;and 3) the cylinder bore acts as a major struc-tural element of the cylinder block, retainingthe cylinderheads and the bottom end com-ponents. The cylinder bore design, finish, andits preparation techniques are extremely im-portant aspects of performance engine de-sign.Cylinder Block —The casting that comprisesthe main structure of an IC engine. The cylin-der block is the connecting unit for the cylin-derheads, crankshaft, and external assem-blies, plus it houses the pistons, camshaftand all other internal engine components. The


stability, strength, and precision of the blockcasting and machining are extremely impor-tant in obtaining optimum power and enginelife. Cylinder blocks are usually made from ahigh grade of cast iron.Cylinderhead —A component (usually madeof cast iron or cast aluminum) that forms thecombustion chambers, intake and exhaustports—including water cooling passages—and provides support for valvetrain compo-nents, spark plugs, intake and exhaust mani-folds, etc. The cylinderhead attaches to theengine block with several large bolts thatsqueeze a head gasket between the blockdeck and head surfaces; and when attached,the head becomes a load-carrying member,adding strength and rigidity to the cylinderblock assembly. Modern cylinderhead designsfall into three major categories: 1) overhead-valve with wedge, canted-valve, or hemi-spherical combustion chambers; 2) single-overhead cam with wedge or hemisphericalcambers; 3) double-overhead cam with hemi-spherical chambers.Degree—1) An angular measurement. A com-plete circle is divided into 360 degrees; equalto one crankshaft rotation; 180 degrees isone-half rotation. 2) A temperature measure-ment. The temperatures of boiling and freez-ing water are: in the Fahrenheit system 212and 32 degrees; in the Celsius system 100and 0 (zero) degrees.Density —A measurement of the amount ofmatter within a known space or volume. Airdensity is the measurement of the amount ofair per unit volume at a fixed temperature,barometric pressure, altitude, etc.Detonation —The secondary ignition of theair/fuel mixture in the combustion space caus-ing extreme pressures. Detonation is causedby low gasoline octane ratings, high combus-tion temperatures, improper combustionchamber shape, too-lean mixtures, etc. Deto-nation produces dangerously high loads onthe engine, and if allowed to continue, willlead to engine failure. Detonation, unlike pre-ignition, requires two simultaneous combus-tion fronts (fuel burning in two or more placesin the combustion chamber at once); whereas

preignition occurs when the fuel-air mix ig-nites (with single burning front) before thespark plug fires. Both preignition and detona-tion produce an audible “knock” or “ping,” butdetonation does not produce the rapid “wildpinging” noise that is typically associated withpreignition. The extreme pressures of deto-nation can lead to preignition, but even worsethe high temperatures of preignition can causedetonation.Duration or Valve Duration —The numberof crankshaft degrees (or much more rarely,camshaft degrees) of rotation that the valvelifter or cam follower is lifted above a speci-fied height; either seat-to-seat valve durationmeasured at 0.006-, 0.010-inch or other valverises (even 0.020-inch lifter rise), or durationmeasured at 0.050-inch lifter rise called 0.050-inch duration. Intake duration is a measureof all the intake lobes and exhaust durationindicates the exhaust timing for all exhaustlobes. Longer cam durations hold the valvesopen longer, often allowing increased cylin-der filling or scavenging at higher enginespeeds.Dynamometer —A device used to measurethe power output of rotating machinery. In it’ssimplest terms, a dynamometer is a power-absorbing brake, incorporating an accuratemethod of measuring how much torque (andhorsepower) is being absorbed. Braking isaccomplished through friction (usually a hy-draulic absorber) or by an electric dynamo(converts energy to electricity). Modern com-puter-controlled dynamometers for high-per-formance automotive use have sophisticatedspeed controls that allow the operator to se-lect the rpm point or range of speeds throughwhich the torque is to be measured. Thenthe operator opens the throttle and the dyna-mometer applies the precise amount of loadto maintain the chosen rpm points; horse-power is read out directly on a gauge and/orcomputer screen.Dynomation —A engine simulation programdeveloped by V.P. Engineering, Inc. ($600;515-986-9197) that uses full wave-actionanalysis, currently using the Method Of Char-acteristics to provide solutions to the com-


Appendix-B Glossary

plex equations of wave dynamics.Empirical or Empirical Testing —Meaning“after the fact” or “experimental,” empiricaltesting involves actual “real-world” experi-ments to determine the outcome of compo-nent changes.Exhaust Center Angle/Centerline or ECA—The distance in crank degrees from the pointof maximum exhaust valve lift (on symmetriccam profiles) to TDC during the valve over-lap period.Exhaust Manifold —An assembly (usually aniron casting) that connects the exhaust portsto the remainder of the exhaust system. Theexhaust manifold may include a heat-riservalve or port that heats the intake manifold toimprove fuel vaporization.Exhaust Ports —Cavities within thecylinderhead that form the initial flow pathsfor the spent gases of combustion. One endof the exhaust port forms the exhaust valveseat and the other end forms a connectingflange to the exhaust manifold or header.Exhaust Stroke —One of the four 180-de-gree full “sweeps” of the piston moving in thecylinder of a four-stroke, internal-combustionengine (originally devised by Nikolaus Otto in1876). During the exhaust stroke, the pistonmoves from BDC to TDC and forces exhaustgases from the cylinder into the exhaust sys-tem. Note: The 180-degree duration of theexhaust stroke is commonly shorter than theperiod during which the exhaust valve is open,sometimes referred to as the true “ExhaustCycle.” The exhaust stroke is followed by theintake stroke.Exhaust Valve Closing or EVC—The pointat which the exhaust valve returns to its seat,or closes. This valve timing point usually oc-curs early in the intake stroke. Although EVCdoes not have substantial effects on engineperformance, it contributes to valve overlap(the termination point of overlap) that canhave a significant effect on engine output.Exhaust Valve Duration —See DurationExhaust Valve Lift —See Valve LiftExhaust Valve Opening or EVO—The pointat which the exhaust valve lifts off of its seat,or opens. This valve timing point usually oc-

curs late in the power stroke. EVO usuallyprecedes BDC on the power stroke to assistexhaust-gas blowdown. This EVO timing pointcan be considered the second most impor-tant cam timing event.Exhaust Valve —The valve located within thecylinderhead that control the flow of spentgases from the cylinder. The exhaust valvesare precisely actuated (opened and closed)by the camshaft, usually through lifters, push-rods, and rockerarms. Exhaust valves mustwithstand extremely high temperatures (1500degrees-F or higher) and are made from spe-cial steels, e.g., SAE J775 that has excellentstrength at high temperatures and good re-sistance to corrosion and wear.Filling & Emptying Multidimensional Simu-lation —This engine simulation technique in-cludes multiple models (e.g., thermodynamic,kinetic, etc.), and by dividing the intake andexhaust passages into a finite series of sec-tions it describes mass flow into and out ofeach section at each degree of crank rota-tion. The Filling And Emptying method canaccurately predict average pressures withinsections of the intake and exhaust systemand dynamically determine VE and enginepower. However, the basic Filling And Emp-tying model can not account for variations inpressure within individual sections due to gasdynamic effects. See Gas-Dynamic Multidi-mensional Simulation.Finite-Amplitude Waves —Pressure wavesof higher energy levels higher than acousticwaves. Finite-amplitude waves exhibit com-plex motions and interactions when travelingthrough engine passages. These actionsmake their mathematical analysis very com-plex.Flat-Tappet Lifter —A camshaft follower hav-ing a flat surface at the point of contact withthe cam lobe. Flat-tappet lifters actually havea shallow convex curvature at their “face” toallow the lifter to rotate during operation, ex-tending the working life.Flow Bench and Flow-Bench Testing —Aflow bench is a testing fixture that develops aprecise pressure differential to either “suck”our “blow” air through a cylinderhead or other


engine component. A flow bench determinesthe flow capacities (restrictions) ofcylinderhead ports and valves and assists inthe analysis of alterations to port contours.Four-Cycle Engine —Originally devised byNikolaus Otto in 1876, the four-cycle engineconsists of a piston moving in a closed cylin-der with two valves (one for inlet and one foroutlet) timed to produce four separate strokes,or functional cycles: Intake, Compression,Power, and Exhaust. Sometimes called the”suck, squeeze, bang, and blow” process, thistechnique—combined with a properly atom-ized air/fuel mixture and a precisely timedspark ignition—produced an engine with highefficiency and power potential. The softwarediscussed in this book is designed to simu-late the functional processes of a four-cycleengine.Friction Horsepower (Fhp) —The power ab-sorbed by the mechanical components of theengine during normal operation. Most frictionallosses are due to piston ring pressure againstthe cylinder walls. Frictional power losses arenot easily measured, however, they can beaccurately calculated knowing the brakehorsepower (from dyno testing) and the indi-cated horsepower (from pressure measure-ments). Also see Indicated Horsepower andBrake Horsepower.Friction Mean Effective Pressure (Fmep) —A theoretical average pressure that wouldhave to be present in each cylinder duringthe power stroke to overcome the power con-sumed by friction within the engine. Fmep isusually calculated by first determining the In-dicated Mean Effective Pressure (Imep)—themaximum horsepower that can be producedfrom the recorded cylinder pressures. TheBrake Mean Effective Pressure (Bmep) is thenmeasured by performing a traditional “dyno”test. The Fmep is calculated by finding thedifference between the Imep and the Bmep:Fmep = Imep - Bmep. Fmep can also bedirectly measured with a motoring (electric)dyno.Friction —A force that opposes motion. Fric-tional forces convert mechanical motion intoheat.

Full Three-Dimensional (CFD) Simulation —This highly advanced engine simulation tech-nique incorporates multiple models (e.g., ther-modynamic, kinetic, etc.), including full three-dimensional modeling that subdivides an area,such as the combustion chamber or port junc-tion, into a series of volumes (or cells) throughwhich the model solves the differential equa-tions of thermodynamics and fluid flow (usingComputational Fluid Dynamics—CFD). Theinteraction of these cells can reveal verysubtle design features within the induction andexhaust systems. It can thoroughly evaluatetheir effect on horsepower, fuel efficiency, andemissions throughout the rpm range. Thissimulation is not available as a commercialprogram and it currently remains a “labora-tory-only” tool.Gas-Dynamic Multidimensional Simula-tion —This engine simulation technique in-cludes multiple models (e.g., thermodynamic,kinetic, etc.), plus powerful finite-wave analy-sis techniques that account for variations inpressures within individual sections of theports due to gas dynamic effects. This de-tailed, highly math-intensive technique canpredict engine power with remarkably highaccuracy. The Dynomation program from V.P.Engineering ($600; 515-986-9197)—dis-cussed in the complete DeskTop Dynosbook—is an example of a program using thissimulation method.Helmholtz Resonator —A device that in-creases the amplitude of pressure wavethrough a resonance phenomenon (an effectsimilar to the deep “whir” produced when airis blown over the neck of a jug). In somecases, the induction system in an IC enginecan be modeled by employing Helmholtzresonance equations.Horsepower —Torque measures how muchwork (an engine) can do, power is the rate-based measurement of how fast the work isbeing done. Starting with the static force ap-plied at the end of a torque arm (torque),then multiplying this force by the swept dis-tance through which the same force wouldrotate one full revolution finds the power perrevolution: Power Per Revolution = Force or


Appendix-B Glossary

Weight x Swept Distance. James Watt (1736-1819) established the current value for onehorsepower: 33,000 pound-feet per minuteor 550 pound-feet per second. So horsepoweris currently calculated as: Horsepower =Power Per Revolution/33,000, which is thesame as Horsepower = (Torque x 2 x Pi xRPM)/33,000, or simply: Horsepower =(Torque x RPM)/5,252. The horsepower be-ing calculated by these equations is just oneof several ways to rate engine power output.Various additional methods for calculating ormeasuring engine horsepower are commonlyused (to derive friction horsepower, indicatedhorsepower, etc.), and each technique pro-vides additional information about the engineunder test.Hydraulic Lifter —See Lifters, HydraulicIC Engine —See Internal Combustion engine.Inches of Mercury and Inches of Water —Astandard method of pressure measurement,where pressures are compared to atmo-spheric or ambient pressure. Inches of dis-placement are recorded for a water or mer-cury column measured in a “U” shaped tubewith one end open to the air and the otherend connected to the test pressure. Com-monly called a manometer, this pressure com-parison device is quite sensitive and accu-rate. When mercury is used in the manom-eter tube, one psi differential from atmosphericpressure will displace 2.04-inches of mercury.However, when water is the liquid in the “U-tube,” a substantial increase in pressure sen-sitivity is obtained: one psi will displace 27.72inches of water. A water manometer is usedto measure small vacuum and pressure sig-nals.Indicated Horsepower (Ihp) —is the maxi-mum power that a particular engine can theo-retically produce. It is calculated from ananalysis of the gas pressures measured byinstalling pressure transducers in the cylin-ders. Also see Brake Horsepower and Fric-tion Horsepower.Indicated Mean Effective Pressure (Imep) —A theoretical average pressure that wouldhave to be present in each cylinder duringthe power stroke to generate the maximum

horsepower possible from the pressures re-corded within the cylinder of an engine dur-ing an actual dyno test. The Imep pressureassumes that the recorded pressures withinthe test engine will be entirely converted intomotive force (with no losses due to friction).Induction Airflow —The airflow rating (ameasurement of restriction) of a carburetoror fuel injection system. Four-barrel carbure-tors are rated by the measured airflow whenthe device is subjected to a pressure dropequal to 1.5-inches of Mercury. Two-barrelcarburetors are tested at 3.0-inches of Mer-cury.Induction System —Consists of the carbure-tor or injection system and the intake mani-fold. The intake manifold can be of many de-signs such as dual plane, single plane, tun-nel ram, etc.Intake Centerline Angle —The distance incrank degrees from the point of maximumintake valve lift (on symmetric cam profiles)to TDC during the valve overlap period.Intake Stroke —One of the four 180-degreefull “sweeps” of the piston moving in the cyl-inder of a four-stroke, internal-combustionengine (originally devised by Nikolaus Otto in1876). During the intake stroke, the pistonmoves from TDC to BDC and inducts (drawsin by lowering the pressure in the cylinder)air/fuel mixture through the induction system.Note: The 180-degree duration of the intakestroke is commonly shorter than the periodduring which the intake valve is open, some-times referred to as the true “Intake Cycle.”The intake stroke is followed by the com-pression stroke.Intake Valve Closing or IVC—Consideredthe most important cam timing event. Thepoint at which the intake valve returns to itsseat, or closes. This valve timing point usu-ally occurs early in the compression stroke.Early IVC helps low-end power by retainingair/fuel mixture in the cylinder and reducingcharge reversion at lower engine speeds. LateIVC increases high-speed performance (at theexpense of low speed power) by allow addi-tional charge to fill the cylinder from the ram-tuning effects of the induction system at higher


engine speeds.Intake Valve Duration —See DurationIntake Valve Lift —See Valve LiftIntake Valve Opening or IVO—The point atwhich the intake valve lifts off of its seat, oropens. This valve timing point usually occurslate in the exhaust stroke. Although IVO doesnot have a substantial effect on engine per-formance, it contributes to valve overlap (thebeginning point of overlap) that can have asignificant effect on engine output.Internal Combustion Engine —An enginethat produces power from the combustion andexpansion of a fuel-and-air mixture within aclosed cylinder. Internal-combustion enginesare based on two methods of operation: twocycle and four cycle. In each method, a mix-ture of fuel and air enters the engine throughthe induction system. A piston compressesthe mixture within a closed cylinder. A pre-cisely timed spark ignites the charge after itis compressed. The explosive burning pro-duces very high temperatures and pressuresthat push the piston down and rotate thecrankshaft, generating a motive force. Alsosee combustion space, compression ratio,compression stroke, power stroke, exhauststroke, and intake stroke.Lifters, Hydraulic Flat-Tappet —A camshaftfollower having a flat surface at the point ofcontact with the cam lobe. Flat-tappet liftersactually have a shallow convex curvature attheir “face” to allow the lifter to rotate duringoperation, extending the working life. A hy-draulic lifter incorporates a mechanism thatautomatically adjusts for small changes incomponent dimensions, and usually maintainszero lash in the valvetrain. Hydraulic liftersalso offer a slight “cushioning” effect and re-duce valvetrain noise.Lifters, Roller Solid Or Hydraulic —A cam-shaft follower having a round, rolling elementused at the contact point with the cam lobe.A hydraulic lifter incorporates a mechanismthat automatically adjusts for small changesin component dimensions, and usually main-tains zero lash in the valvetrain. Hydrauliclifters also offer a slight “cushioning” effectand reduce valvetrain noise. Solid lifters lack

this hydraulic adjusting mechanism and re-quire a running clearance in the valvetrain,usually adjusted by a screw or nut on therockerarm.Lifters, Solid Flat-Tappet — A camshaft fol-lower having a flat surface at the point ofcontact with the cam lobe. Flat-tappet liftersactually have a shallow convex curvature attheir “face” to allow the lifter to rotate duringoperation, extending the working life. Solidlifters lack an automatic hydraulic adjustingmechanism and require a running clearancein the valvetrain, usually adjusted by a screwor nut on the rockerarm. Solid lifter cams usu-ally generate more valvetrain noise than hy-draulic-tappet cam.Lobe-Center Angle or LCA—The angle incam degrees from maximum intake lift tomaximum exhaust lift. Typical LCAs rangefrom 100 to 116 camshaft degrees (or 200 to232 crank degrees).Multi Dimensional —As it refers to enginesimulation programs, multi dimensional indi-cates that the simulation is based on multiplemodels, such as thermodynamic and kinetic,plus the multidimensional geometric descrip-tion of inlet and outlet passages and a dy-namic model of induction and exhaust flow.Negative Pressure —A pressure below at-mospheric pressure; below 14.7psi absolute.Very low pressures are usually measured bya manometer. See Inches of Water.Normally Aspirated —When the air-fuel mixis inducted into the engine solely by the lowerpressure produced in the cylinder during theintake stroke; aspiration not aided by a su-percharger.Otto-Cycle Engine —See Four-Cycle EngineOverlap or Valve Overlap —The period, mea-sured in crank degrees, when both the ex-haust valve and the intake valve are open.Valve overlap allows the negative pressurescavenge wave to return from the exhaustand begin the inflow of air/fuel mixture intothe cylinder even before the intake strokebegins. The effectiveness of the overlap pe-riod is dependent on engine speed and ex-haust “tuning.”Pocket Porting— Relatively minor porting


Appendix-B Glossary

work performed below the valve seat and inthe “bowl” area under the valve head. Thesechanges, while straightforward, can producea significant improvement in airflow and per-formance. Proper contours must be main-tained, particularly below the valve seat, toproduce the desired results.Porting or All-Out Porting —Aggressive port-ing work performed to the passages withinthe cylinderhead with intention of optimizinghigh-speed airflow. Often characterized bylarge cross-sectional port areas, these portsgenerate sufficient flow velocities only athigher engine speeds; low speeds produceweak ram-tuning effects and exhaust scav-enging waves. This porting technique is apoor choice for low-speed power and streetapplications.Pounds Per Square Inch— See PSI.Power Stroke —One of the four 180-degreefull “sweeps” of the piston moving in the cyl-inder of a four-stroke, internal-combustionengine (originally devised by Nikolaus Otto in1876). During the power stroke, the pistonmoves from TDC to BDC as the burning air/fuel mixture forces the piston down the cylin-der. Note: The 180-degree duration of thepower stroke is commonly longer than theduration between top dead center and theexhaust-valve opening point, sometimes re-ferred to as the true “Power Cycle.” The powerstroke is followed by the exhaust stroke.Pressure Crank-Angle Diagram —Alsocalled a “Pie-Theta ”diagram, the crank-anglediagram is a plot of the indicated cylinderpressures vs. the angular position of thecrankshaft during the entire four-cycle pro-cess. This diagram provides an easily under-stood view of the varying pressures in thecylinder. Also see Pressure-Volume Diagram.Pressure —A force applied to a specificamount of surface area. A common unit ofpressure is psi, i.e., pounds per square inch.The force that develops the pressure is sumtotal of all the slight “nudges” on a surfacegenerated by each molecule striking the sur-face; the greater number of impacts or themore violent each impact, the greater thepressure. Therefore, the pressure increases

if the same number of molecules are con-tained in a smaller space (greater number ofimpacts per unit area) or if the molecules areheated (each impact is more violent).Pressure-Volume Diagram —Also called aPV (pronounced Pee-Vee), or “indicator” dia-gram, the pressure-volume diagram plots in-dicated pressure against the displaced vol-ume in the cylinder. A PV diagram has theremarkable feature of isolating the work con-sumed from the work developed by the en-gine. The area within the lower loop, drawnin a counterclockwise direction, represents thework consumed by the engine “pumping” thecharge into the cylinder and forcing the ex-haust gasses from the cylinder. The upperloop area, drawn in a clockwise direction, in-dicates the work produced by the engine frompressures generated after combustion.PSI or Pounds Per Square Inch —A mea-sure of force applied on a surface, e.g., theforce on the wall of a cylinder that contains acompressed gas. A gas compressed to 100psi would generate a force of 100 pounds oneach square inch of the cylinder wall surface.In other words, psi equals the force in poundsdivided by the surface area in square inches.Also see pressure.Restriction —A measure of the resistance toflow for (usually) a liquid or gas. Exhaust orintake flow restriction can occur within tubingbends, within ports, manifolds, etc. Liquid re-striction can occur in needle-and-seat assem-blies, fuel pumps, etc. Some restriction is al-ways present in a flowing medium.Roller Lifter —See Lifters, RollerRoller Tappet Lifter — See Lifters, RollerRPM—Revolutions Per Minute. A unit of mea-sure for angular speed. As applied to the ICengine, rpm indicates the instantaneous ro-tational speed of the crankshaft described asthe number of crank revolutions that wouldoccur every minute if that instantaneousspeed was held constant throughout the mea-surement period. Typical idle speeds are 300to 800rpm, while peak engine speeds canreach as high as 10,000rpm or higher in someracing engines.Seat-To-Seat Cam Timing Method —See


Cam Timing, @ Seat to Seat.Simulation and Engine Simulation —A en-gine simulation process or program attemptsto predict real-world responses from specificcomponent assemblies by applying funda-mental physical laws to “duplicate” or simu-late the processes taking place within thecomponents.Smallblock —A generic term that usually re-fers to a V8 engine with a displacement smallenough to be contained within a “small” sizeengine block. Typical smallblock engines dis-place under 400 cubic inches.Solid Lifter —See Lifter, SolidStroke —The maximum distance the pistontravels from the top of the cylinder (at TDC)to the bottom of the cylinder (at BDC), mea-sured in inches or millimeters. The stroke isdetermined by the design of the crankshaft(the length of the stroke arm).Top Dead Center or TDC—The position ofthe piston in the cylinder bore at its upper-most point in the stroke. Occurs twice withinthe full cycle of a four-stroke engine; at thestart of the intake stroke and 360 degreeslater at the end of the compression stroke.Torque —The static twisting force producedby an engine. Torque varies with the lengthof the “arm” at which the twisting force ismeasured. Torque is a force times the lengthof the measurement arm: Torque = Force xTorque Arm, where Force is the applied orthe generated force and Torque Arm is thelength through which that force is applied.Typical torque values are ounce-inches,pound-feet, etc.Valve Head and Valve Diameter —The largeend of an intake or exhaust valve that deter-mines the diameter. Valve head temperaturecan exceed 1200 degrees F during engineoperation and a great deal of that heat istransferred to the cylinderhead through thecontact surface between the valve face andvalve seat.Valve Lift Rate —A measurement of how fast(in inches/degree) the camshaft raises thevalve off of the valve seat to a specific height.If maximum valve lift is increased but the du-ration (crankshaft degrees) that the valve is

held off of the valve seat is kept the same,then rate at which the valve opens must in-crease (same time to reach a higher lift). Highlift rates can produce more horsepower, how-ever, they also increase stress and valve-train wear.Valve Lift —The distance the valve headraises off of the valve seat as it is actuatedthrough the valvetrain by the camshaft. Maxi-mum valve lift is the greatest height the valvehead moves off of the valve seat; it is the liftof the cam (lobe height minus base-circle di-ameter) multiplied by the rockerarm ratio.Valve Motion Curve or Valve DisplacementCurve —The movement (or lift) of the valverelative to the position of the crankshaft. Dif-ferent cam styles (i.e., flat, mushroom, orroller) typically have different displacementcurve acceleration rates. Engine simulationprograms calculate a valve motion curve fromvalve event timing, maximum valve lift, andother cam timing specifications.Volumetric Efficiency —Is calculated by di-viding the mass of air inducted into the cylin-der between IVO and IVC divided by the massof air that would fill the cylinder at atmosphericpressure (with the piston at BDC). Typicalvalues range from 0.6 to 1.2, or 60% to 120%.Peak torque always occurs at the enginespeed that produced the highest volumetricefficiency.Work and Net Work —Work is the energyrequired to move an object over a set dis-tance. Both motion and force must be presentfor work to occur. In an IC engine, work isdeveloped from pressures within the cylinderacting on the face of the piston (producing aforce) times the distance through which thepiston travels. In the internal combustion en-gine some of the work produces power out-put (such as pressures producing pistonmovement during the power stroke) and someof the work is negative (like compressing thefresh charge on the compression stroke). Thedifference between the positive and negativework is the net work produced by the engine.


Anderson, Edwin P., Facklam,Charles G, Gas Engine Manual ,G.K. Hall & Co., 70 Lincoln Street,Boston, MA 02111, 800-343-2806. Avery readable, non-engineering lookat the four and two-cycle internalcombustion engine. Includes a de-tailed look at individual componentsand their function. Published 1962.ISBN-0-8161-1707-1

Cummins Jr., C. Lyle, Internal Fire ,Internal Fire, SAE. In-depth historyof the internal-combustion engine,from the early attempts at gunpow-der engines in the 1600’s to “mod-ern” designs of the early 1900’s.Easy, interesting reading. Orig. Pub-lished 1976, revised 1989. ISBN 0-89883-765-0

Benson, Roland S., The Thermody-namics and Gas Dynamics of In-ternal-Combustion Engines, Vol-ume 1 , Oxford University Press, NewYork. Considered by many to be theoriginal “bible” of gas dynamics asapplied to the IC engine. This bookhas been the jumping off point formany engineers and scientists intothe world of computer engine model-ing. This book is mathematically rig-orous and demands much of the

The following books are generally not available in auto stores or “speed shops.” Thesereferences can be found in well-stocked engineering libraries (particularly at state uni-versities) or at technical book stores. Many of these books are serious engineeringworks requiring substantial math and physics background for complete understand-ing, however, all of these books have least some “readable” material that might be ofinterest to performance enthusiasts.

reader. A knowledge of differentialand integral calculus is required toget the most from this text. Published1982 for the late Rowland Benson(1925-1978) by J.H. Horlock F.R.Sand D.E. Winterbone, editors. ISBN0-19-856210-1

Ferguson, Colin R., Internal Com-bustion Engines, AppliedThermosciences , John Wiley &Sons, New York, An interesting com-bination of “nuts and bolts” and ther-modynamics. A much more rigorouslook at thermal science (encom-passes two thirds of the book) thanthe design of IC engines. Requiresknowledge of differential and integralcalculus. Includes some computercode listings, however, calls aremade to subroutines that may not beavailable to many readers. Published1986. ISBN 0-471-88129-5

Stone, Richard, Introduction To In-ternal Combustion Engines ,Macmillan Publishing. A text for col-lege-level students. Includes somehigh-level math, but much of the textis accessible to the advanced enthu-siast. Covers thermodynamics (witha quick look at modeling theory),combustion, turbocharging, mechani-

Appendix-C Bibliography


cal design, dynamometers. Published1985; second edition published 1992.ISBN 0-333-37593-9

Heywood, John B., Internal Com-bustion Engine Fundamentals ,McGraw-Hill, Inc. A very thorough textfor researchers and professionals.Includes both in-depth technical treat-ments (with thorough mathematicalanalysis) and a considerable amountof material that should be accessibleto many enthusiasts. Nearly 1000pages with many fascinating photosand drawings of research work. Pub-lished 1988. ISBN 0-07-028637-X

Markatos, N.C., Computer Simula-tion For Fluid Flow, Heat And MassTransfer, And Combustion In Re-ciprocating Engines , HemispherePublishing Corp., New York. A bookthat evolved from a course on com-puter simulation for fluid flow held inDubrovnik, Yugoslavia in September1987. Heavy mathematical treatment;very little light reading here. Published1989. ISBN 0-89116-392-1

Ramos, J.I., Internal CombustionEngine Modeling , Hemisphere Pub-lishing Corp., New York. A text in-tended for researchers in fluid me-chanics, combustion, turbulence andheat transfer at the graduate level.Rigorous mathematical treatment in-cludes rotary, diesel, two- and four-stroke spark ignition engines. Pub-

lished 1989. ISBN 0-89116-157-0

Ganesan, V., Internal CombustionEngines , Tata McGraw-Hill Publish-ing Company Limited. A book writtenby V. Ganesan, a Professor of Me-chanical Engineering at the IndianInstitute of Technology in Madras,India. Readable text intended for thestudent beginning serious study of theI.C. engine. Good balance betweentextual explanations and mathemati-cal/physics study. Contains manysolved problems. First published1994. ISBN 0-07-462122-X


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ENGINE & DYNO BASICSTHE OTTO-CYCLE ENGINETHE OTTO-CYCLE PROCESSMODERN ENGINE DEVELOPMENTTHE DYNAMOMETERUNDERSTANDING TORQUEUNDERSTANDING POWERUNDERSTANDING HORSEPOWERHORSEPOWER BY ANY OTHER NAMEREAL WORLD HORSEPOWER & TORQUEOTHER TYPES OF DYNOSTHE HYDRAULIC DYNAMOMETERTHE ELECTRIC DYNAMOMETERCOURSE OF DEVELOPMENTTHE ULTIMATE TRIAL-AND-ERROR TOOL

SIMULATING ENGINE PRESSURESENGINE PRESSURESINDICATED CYLINDER PRESSURESTHE PRESSURE CRANK-ANGLE DIAGRAMTHE PRESSURE VOLUME DIAGRAMCALCULATING HORSEPOWERCALCULATING CYLINDER PRESSURESINTAKE CYCLE PRESSURESPOWER CYCLE PRESSURES

EXHAUST CYCLE PRESSURESINDICATED HORSEPOWER AND MEPSIMULATING FRICTIONAL LOSSES

THE GROWTH OF SIMULATIONSEARLY MODELING EFFORTSZERO- AND QUASI-DIMENSIONAL MODELSMULTIDIMENSIONAL MODELSFILLING-AND-EMPTYING MODELSGAS-DYNAMIC MODELSMODERN ENGINE MODELING AND THE PCDYNOMATION:GLEAM IN A YOUNG MAN’S EYETHE FUTURE OF IC MODELING

GAS DYNAMICS AND DYNOMATIONIC ENGINE:AN UNSTEADY FLOW MACHINEACOUSTIC VS. FINITE-AMPLITUDE WAVESCOMPRESSION AND EXPANSION WAVESPRESSURE WAVES AND ENGINE TUNINGPRESSURE-TIME HISTORIESGAS FLOW VS. ENGINE PRESSURESINTAKE TUNINGINDUCTION RUNNER TAPER ANGLESPORT FLOW VELOCITIESEXHAUST TUNINGEXHAUST FLOW VELOCITIESVALVE EVENTS AND STRATEGIESOPTIMIZING VALVE EVENTS

This book picks up where the MiniGuide leaves off. The complete Desk-Top Dynos book will reveal more to youabout high-performance and racing en-gines than you’ve discovered in all otherenthusiast books and magazines com-bined! Typical engine books describe an“outside-in” approach—the direct resultof trial-and-error testing. This book fol-lows a new path. It describes enginefunction from the “inside-out,” explain-ing WHY engines “need” specific partsand WHICH tuning techniques optimize power. DeskTop Dynos clarifies anddistills the facts to give you a deeper understanding of the internal-combus-tion engine for street, high-performance, or racing applications.


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