TECH-LINE

TERMS & DEFINITIONS

Racing & Engine Building have many terms and definitions. The following are some of the most common and often confusing terms used, which can make even a seasoned engine builder pause in thought. This page was originally intended to be only a summary of information explaining some of the terms we use in valve trains, which by itself can be quite extensive. But in adding this information I found that more questions came up which were associated to cylinder heads and overall engine building. Because valve train is an integral part of cylinder head design, and because we address so many problems of valve trains from missed information or incorrect assumptions about cylinder heads, I've expanded this section to include more depth on both of these categories. I've also included some basics on Turbos, and added comments and links on the major MEI products. Lastly, a few rudimentary terms and definitions have been included as an introductory courtesy to the novice engine builders just starting out. We've all been there! Good luck.

--jM

1/3 RULE: Refers to the traditional method engineers used for rocker arm design on an Over-Head Valve engine. See GEOMETRY, TRADITIONAL.

ABDC: Cam lingo, for AFTER BDC (Bottom Dead Center). Sometimes used on CAM CARDS.

ATDC: Cam lingo, for AFTER TDC (Top Dead Center). Sometimes used on CAM CARDS.

ACCELERATION DYNAMICS, VALVE: The "net" valve lift, acceleration rate and duration (time open); or what the valve actually sees as a result of the camshaft's instructions - as translated by the rocker arm design geometry and it's installed geometry. Since a cam profile is actually a blueprint for accelerating the valve train from closed, to open, then trying to slow it down on the closing side, acceleration "dynamics" is changed by the rocker arm's geometry on either end. So when we add this term to the valve, we are really talking about the effect that rocker ratio and geometry have upon the velocity imposed at the valve, from what the cam's acceleration initiated. See: VALVE LIFT DYNAMICS (VLD).

ACCELERATION RAMP: (Camshaft) Refers to the specific portion of a CAM LOBE that lies between the BASE CIRCLE and the CAM'S NOSE, which is precisely calculated to increase the TAPPET lift at a specific rate of acceleration, then slowing as the tappet rides across the cam's peak lift point, and begins to following the DECELERATION RAMP, where the compressed pressure, inertia and mass of the entire valve train are hopefully constrained within operating limits of the desired RPM's, to allow the valve to seat with minimum bounce.

ADJUSTING NUT (ADJUSTER): Traditionally named on STUD MOUNT rocker arms for securing the rocker arm upon the rocker's mounting stud. Incorporates a SET SCREW which locks using an Allen wrench, to secure more tightly than OEM factory interference fit adjusting nuts when setting valve lash on performance camshafts.

ADJUSTING SCREW: Self descriptive, usually 3/8" or 7/16" SAE (fine) thread, with either a 3/8" diameter "ball" or a 5/16" diameter "cup" on the mating end for pushrod contact; used to facilitate valve lash adjustment on fixed height, stand mounted rocker arms (also known as shaft systems), or non-stud type mounting.

AFTERCOOLER: (Turbocharger) Often mislabeled as an "Intercooler," even by manufacturers and distributors of Turbochargers and accessories, the Aftercooler is essentially a radiator, designed in one of two fashions, whereby it reduces the hot air temperature exiting the COMPRESSOR side of the Turbocharger, before being sent to the engine's induction system. Cylinder operating temperatures and fuel burn performance are more difficult with the warmer air, and they reduce the density of the air fuel mixture, and thus the engine's power. Coming in many shapes and sizes, the two basic design approaches of Aftercoolers is "air-to-air" and "liquid" cooled. In each case, the induction system air being forced into the engine from the Turbocharger's Compressor is routed through this aluminum "Heat Exchanger" that is comprised of a grid of small passages separated by, and attached to dissipating fins, just like a common radiator, whereby either (1) AIR directly passes across them for cooling, or (2) Liquid (usually water) flows across in carrying off the radiant heat. In the air cooled version, the Aftercooler must be designed and installed in such a way that the air passing through is forced and as cool as ambient conditions will allow, and of course separated from other nearby heat sources from the engine and/or Turbocharger. Sometimes a difficult task in tightly fitted engine bays. Water cooling versions, usually more complex, offer more latitude with such restrictions, but at the expense of more weight, so they are less common. Manifold tubing from the Compressor to and from the Aftercooler, should be designed with the same efficiency of airflow used on any induction system, with minimal lengths and bends being first goals. The efficiency of Aftercoolers pertains to how much reduction of the Compressor Discharge Temperature (CDT) can be attained, and in most cases the best of them will be in the 60% to 65% range. Meaning: whatever the CDT temperature is over the ambient Compressor Inlet Temperature (CIT) , approximately 60% of that difference can be reduced. This post-Aftercooler Temperature is known as the Air Inlet Temperature (AIT), which is measured within the manifold runner between the Aftercooler and the throttle body of the engine, but usually as close to the throttle body inlet as possible.

NOTE: Because CDT temperatures which often exceed 200° Fahrenheit make less power than cooler induction system temperatures that can use more fuel. These higher induction temperatures elevate the threshold  for controlling detonation in the combustion chamber. Although the "boost" (and air flow) lost through the airflow's transition from one end to the other of a good, properly sized Aftercooler is usually about 2" of Mercury, most tests show the benefits from the cooler air alone outweigh this. Especially when many systems can be easily adjusted to compensate.

AIR INLET TEMPERATURE (AIT): (Turbochargers) Sometimes confused as the temperature of air entering the Turbocharger, it is really the air entering the engine at the throttlebody. It is particularly the result of cooling the much warmer CDT (Compressor Discharge Temperature) air exiting the Aftercooler, on Turbocharged engines so equipped.

 ALPHA BAR : The cool looking horizontal bar of "alphabetical" letters at the TOP  ^  of THIS LIST, that YOU should be using to navigate this page. Choose the FIRST LETTER of the word GROUP that you're looking for, the scroll down until you hopefully find it. You can E-MAIL us on any words you'd like to see here.

NOTE: HIGHLIGHTED, off-color words in any term will take you directly to that definition automatically. All UNDERLINED RED words are active Hyperlinks that take you to a new web page (except this example).

ANGLES: Is the fundamental division of a circle or any derivative thereof, whereby the measurement stems from a common axis or tangent point. A slice of pie's axis is the point where two edges come together. These measurements are referred to in DEGREES, MINUTES and SECONDS. Degrees is the full measure of a circle, and is divided into the equal numerical value of 360, while MINUTES and SECONDS are subsequent divisions having a value (as with a clock) of 60 for each degree measured. The symbols for this type of measurement is an apostrophe for MINUTES and quotation marks for SECONDS, and looks like this: 11º 20' 00". This is Chevy's 23º Small Block Head rocker STUD angle to the valve. Since the 20 minutes is even, the seconds value is "00."  ^

A second value system for measuring the fractional values of one degree, (in lieu of minutes and seconds) is DECIMAL. In this second alternative to MINUTES and SECONDS, a single degree is divided by 10th's, 100th's, 1000th's and so on. There is no division of the decimal measurement of one degree, merely the finer tolerances of more decimal locations. Our Chevy Stud angle measured like this would look like the following in decimal: 11.3333º. Depending on how many decimal placements are used (10th's or 1,000th's), the more accurate terminology (and ORIGINAL) standard is still "Degrees, Minutes and Seconds," unless going to the fourth decimal place. But this is not usually done. When converting between one or the other, understand that there is no higher value of MINUTES or SECONDS than "60." So a decimal value of "0.5" degrees is the equivalent of "30 minutes," or another way to see the comparative would be "45 minutes" equals "0.75" degrees. NOTE that the "0.5" and "0.75" both have the "0" in front of their decimal, to be sure there is no confusion that you are talking about a division of LESS than one degree. A small block Chevrolet's ROCKER STUD leans into the valve 11 degrees 20 minutes. The decimal equivalent of this is 11.333 degrees; because 20 minutes is 1/3 of 60, and 1/3 of decimal values (based on 10ths and 100's) is going to be our "0.333" added to the 11 degrees. The traditional system of measuring angles in their full depth of Degrees, Minutes and Seconds, can get very specific. The ultimate division of one degree is found by first dividing by 60, then dividing each of the minutes by 60 again. This is the  same as multiplying 60 X 60, which is 3,600. So you can divide one degree into 3,600 equal parts, by using the traditional system. To get a finer measurement than this using decimal, you would have to take a decimal value out to the FOURTH decimal. Because 1/3600 is equal to "0.0002777". Which is a very tight tolerance when you're measuring angles. The DIGITAL PROTRACTORS we often use and sell for engine builders are limited to 1/10th Degree, compared to the more precise Protractors we use from Mahr, a German gauge manufacturer which cuts degrees into minutes and seconds. But the Digital Protractors of 1/10th degree will still be infinitely more accurate than anything any engine builder will ever require. Case in point: Our small block Chevrolet rocker arm will move about .023" for each degree of rotation. If you are using our Digital protractor atop your rocker arm to see what's happening in "angles" instead of linear measurements; then for each 1/10 indication of the Protractor, you've just witnessed .0023" (TWO and THREE TEN THOUSANDTHS of an inch) motion. Is that accurate... or what? (See: TOOLS)  ^

An unrelated point of interest: As the pilot of a vessel that flies or floats, you soon need to learn Nautical navigation, or at least some of its fundamentals. Among those who have, it is common knowledge that a Statute mile is 5,280 feet, and a Nautical mile is 6,072 feet. But one interesting thing that many navigators don't know about the history of this measurement, is where the actual dimension of a nautical mile comes from. They simply accept that a statute mile 1.15 times greater than a nautical mile when they are converting for this. The earth is really cut up in measurements using degrees and angles much in the same division of measurement engineering uses. Even though it's not germane to terminologies of rocker arms and engines, there is a common element to this engineering method which has parallels. The very essence of a Nautical mile's length is based upon the equal division of the earth's diameter at the Equator. If you look at a nautical chart, or a simple world map (which often uses nautical references), you'll see our angular measurement shown in DEGREES, MINUTES and SECONDS. What nautical navigators learn is that the MINUTE value of navigation is equal to a Nautical mile. But what few realize until they think about it, is that the 6,072 feet was derived from slicing the earth's diameter at the equator from a 24 hour rotation into 21,600 equal lengths (360 degrees multiplied times 60 minutes); and this distance turned out to be our 6,072 feet, per "piece," or our Nautical mile. The same value is given from North Pole to South Pole (Longitude), the Latitudinal lines in each direction from the Equator progressively show the minute values narrowing as they approach their apex. So the accuracy on a map for the minute "clicks" (as they are known) is only relevant at the EQUATOR and along the Longitudinal lines (North to South). You'll see both conventional and digital terms of angular measurement used with GLOBAL POSITIONING SYSTEMS (GPS), now found in our everyday aspects of travel. When dealing with something as big as the Earth, you can get more accuracy from a GPS by setting them to DECIMAL. If you set a GPS to the 5th decimal, you can actually cut the earth into roughly 6 foot spaces! But all this isn't relevant here, except to point out that at the core of many things, a common standard exists, and sometimes it is interesting (if not beneficial) to understand what these common denominators are.  ^

ANGLE FORMULA: (MILLER MID-LIFT ROCKER ARM PRINCIPLES) We use a mean sine value of .0173"/per inch/per degree to calculate rocker radial motion (when we don't have our computer handy). You can determine valve lift from degrees of motion, or degrees of motion from valve lift, or degrees of motion from cam lift. You multiply the pivot length (either the pushrod side or the valve side, depending on your goal) times a mean sine value of .0173"/per inch/per degree. The sum is then divided into either the cam lift or the valve lift, to determine the degrees of motion. If your quest is measured from the cam lift side, you must know the distance from the trunnion to the center of the pushrod cup. Multiply that figure times .0173" will give you how many thousandths of an inch motion will occur for EACH degree. If you need to know how many degrees of rotation the rocker will move, you merely need to divide your net cam lift by this figure. If your quest pertains to the valve side, then you need to know the distance from the trunnion to the roller pin centerline, and multiply that times .0173". If you already determined the rocker's rotational value in the pushrod side I just explained, then now you can determine what your net valve lift "should be," by multiplying the degrees of rotation times the sum of your rocker's valve side pivot length (roller tip center-to-center to trunnion). Don't be surprised if you find something not matching between "should" and "reality" with everyone else's rocker arms. Only MID-LIFT Geometry makes these values predictable and accurate. But remember, all rocker arms will have a "flex" compensation factor built in, if they are really accurate on net ratio. Including MILLER MID-LIFT®. The difference though, is we need "less" than everyone else, because they are correcting for geometry errors too; and thus, losing precious area-under-the-curve! Cosine formulas are a constantly cumulative value based on each increase in angle, but you will find that .0173" inches, per degree, per inch of motion is a very close mean (average). Closer than you'll ever use on an engine.  ^

EXAMPLE: Trunnion to Push-rod Cup Center-to-Center: .840"; CAM LIFT: .400". Multiply .0173" X .840" = .0145". To attain DEGREES of ROTATION: Divide .400" (CAM lift) by .0145= 27.58° (decimal); or 27°34'.

EXAMPLE: Trunnion to Roller Tip Center-to-Center: 1.560"; VALVE LIFT: .750". Multiply .0173" X 1.560" = .0269". To attain DEGREES of ROTATION: Divide .750" (Valve lift) by .0269= 27.88° (decimal); or 27°52'.

ANGULAR ACCELERATION: Refers to the rate of change in the velocity of a device rotating around an axis, as measured from e specific perspective. In our sector (engines) this would pertain to the "perspective" of the valve centerline and the device would of course be the rocker arm. This term is sometimes associated with lengths of a device in comparison to another; such as a rocker arm that is longer over another. The term might be used to describe the degree of reduced angular motion of a longer rocker to impose the same net valve lift. The angle of motion for a longer instrument is less to impart the same amount of valve lift.

NOTE: This observation is argued as beneficial to bearing reciprocation speed (since it is rotating less degrees), in and out arc at the rocker's end (because the rocker length/radius is greater), and the inertia values are influenced. These arguments for purposely extending a rocker arm's length are dismissed by one simple value, "aspect ratio." Because to make value of the longer rocker argument, you need to add more mass, which is more weight; and there is also more flex, because you are still asking a device with a similar HEIGHT (bearing bore and bearing bore housing area) to remain the same (to fit the head and stand), while you are extending its length to appease this failed argument of "longer is better." The ASPECT ratio of rocker body mass is now STRETCHED in the wrong direction (greater aspect ratio), and all these other detriments have now come into play: more weight, mass, inertia, flex. The MOMENT of INERTIA has now been extended further from its axial point (shaft centerline). None of which are a virtue to this argument. Next time you see the most advertised shaft rocker manufacturer arguing for his longer rocker theories, think about the above. There is only ONE TIME when you want to stretch a rocker arm beyond what is necessary for hard part clearances to the rocker body (valve springs, retainers, etc.). We do this under a Patent Pending concept with the PRO-SHAFT and PRO-STAND systems.  ^

AREA-UNDER-THE-CURVE: is a term which describes what a graph of the valve lift cycle would look like if charted on paper horizontally across TIME, in crank degrees of rotation, and vertically in thousands of an inch of LIFT. The quicker the valve opens, and "hangs" there, the greater the space beneath the curve from opening to closing. This "area" is what all cam and engine builders seek, because it maximizes cam efficiency. Unfortunately, the recording and studying of VALVE LIFT DYNAMICS (VLD) is rarely practiced by even the best of engine builders, until they learn how important this perspective is. Partly ignored because their previous tests are now inconclusive because there was no diligent record keeping of the VLD. Without practicing correct "INSTALLED Geometry," it is never fully realized at the valve, Of course, you can't have this with a rocker arm that is NOT MID-LIFT in "DESIGN Geometry" either. You need BOTH. MID-LIFT rocker arms deliver this, but need to be installed correctly, which is why we've always included reference planes on the rocker arm which facilitate this for the engine builder.

NOTE: Neither of the measurement references used for cams (lift and duration) represent proper rocker design or geometry when checking NET RATIO. Even the acceleration varies at the valve based on rocker geometry. Therefore, just because you measure another rocker system and find the NET valve lift to be close (or dead on) to the advertised value (based on the cam times ratio formula), doesn't mean the maximum ratio with the least deviation is set throughout the rest of the valve lift cycle. That is because of the adjusting screw's position in the rocker. Traditionally, all other systems have had this too high, and when they didn't achieve accurate NET ratio, in later years (about 1995 and up), better known companies began moving the dimension in closer to the shaft (or trunnion) and placing a correction value into the ratio. The original problem of the adjusting screw (or pushrod cup) being in the wrong place and at the wrong angle was still there. So the rocker was slow coming off the early opening and trailed behind the possible velocity it could have had, until reaching it's full valve lift. The exaggerated ratio, caused by this excessive correction factor on the mathematical "paper" dimensions for where the adjusting screw was at, gave everyone their accurate "net" ratio. But because the Design geometry was not accurate to the pushrod angles, the pushrod was leaving its linear path, wrapping around the rocker axis and thus being slow to lift the valve as quickly as it could have.

Area-under-the-curve, can be seen, measured and compared by graphing out the acceleration rate of the VALVE on a piece of paper. Use graphing paper, and mark the vertical lines from left to right along the bottom of the graph paper to represent CRANK DEGREES of DURATION. Mark the horizontal lines from bottom to top along the left side of the paper to represent VALVE LIFT.
There are TWO perspectives you can use, both achieve the same thing.
1. Select an even spacing of CRANK degrees, like every 5 or 10, and turn the crank over in these even amounts to place a dot where the NET valve lift is on your graph, or...
2. Select an even VALVE lift spacing, like every .010" or .020" and rotate the crank to open the valves, while noting the amount of crank degrees it takes to reach whatever your chosen spacing points is.
You'll get the same result either way, and the amount of gap between your chosen spacing is going to determine how specific and accurate your measurement is going to be; but it also takes more time with the finer jumps. Some people have Cam Doctors, and of course this makes life easy.
Once you "mapped" a particular setup, regardless of who's rockers you're checking; then do the same thing again with a .100" different length pushrod. Doesn't matter which way you go, longer or shorter, just make the change and repeat the process. You'll see the effect AT THE VALVE (where it counts). You may also notice that there was very little change in NET valve lift, and ratio remained very close, in spite of this other loss at the lower valve lifts. What will you see? You'll see a significant difference in lost valve lift for the same amount of crank rotation; or put another way, you'll see it takes the crank more degrees of rotation for the same amount of valve lift. That's the lost area-under-the curve (valve lift "curve") that is being talked about, but rarely understood and even more often wrongly explained by both cam and rocker manufacturers.

ASPECT RATIO: This is a term used in a variety of forums, from engineering and physics to art. It simply means the divided sum of a dimension by an associated second dimension. I've used this term to describe the reality of support a valve guide's length gives to a valve stem in comparison to the off-center forces argued by some that are associated with the rocker arm not being centered in the middle of the valve. In this example, I cite the 2.500" LENGTH of a valve guide supporting a WIDTH of only .342", so the .342" is divided into the 2.500" for an "aspect ratio" value of 7.3. That's over SEVEN times greater support than the force applied to the width of this support. The main point made in this example is that it is not the off-center contact of a rocker's tip atop the valve that is critical (as long as it is entirely on the top of the valve), as much as it is the ANGLE of LOAD that is imposed by the rocker arm's arcing motion.

ASSEMBLY LENGTHS: Another big factor in determining your pushrod length involves any changes made to the head deck, valve seat depth, valve stem length, the block deck, and the cam's base circle. Different lifter heights must also be considered; for instance, you can NOT use the same pushrods for a roller cam, as that which was used for a mechanical or hydraulic cam - since the vast majority of roller lifters are much higher than typical lifters.  ^

ASYMMETRICAL: The two sides of a concept or design which do NOT mirror each other; each having a unique profile, however still dependent upon each other for completing a task. The opposite of symmetrical. See: SYMMETRICAL. Asymmetrical and Symmetrical are two terms often used in CAMSHAFT terminology, referring to the simpler, older cam profiles which had even open and closing ramps, or the more current and most frequently used since the mid-1970's, asymmetrical profiles which lift the tappet quickly to open, and use slower deceleration ramps to set the tappet down in closing. See: CAM RAMPS.

ATTACK ANGLE: Refers to inclines toward or away from operating linear components, like the pushrod's incline towards or away from the valve centerline, as it operates on an assembled engine. We also refer to this for the mounting stud or rocker pad's incline with the valve, which highly influences the range of valve lengths to be used, or rocker lengths, as measured from the axis of rotation to the roller's centerline. In CYLINDER HEADS we use this term to define the PORT WINDOWS created with various port entrance and exit angles to and from the COMBUSTION CHAMBER. See: ROCKER MOUNTING & VALVE TIP HEIGHT for illustrations that depict port silhouettes within the head, relative to valves.

AXIS: The true center of any shaft or pin device in which a another device (usually bearing or roller) rotates about.

BDC: BOTTOM DEAD CENTER: Refers to the position of the PISTON (and crankshaft) in relation to the stroke of the crank. It is the ultimate bottom of the crankshaft's stroke, before initiating its travel back up.

BBDC: Cam lingo, for BEFORE BDC (Bottom Dead Center). Sometimes used on CAM CARDS.

BTDC: Cam lingo, for BEFORE TDC (Top Dead Center). Sometimes used on CAM CARDS.

BACK CUT: (Cylinder Heads) Refers to an angle that is precision ground on the valve and behind the VALVE FACE angle, which is always a lesser value, and which is specifically intended to do two things: First, to adjust and reduce the width of the Valve Face angle to a specific width, usually equal (or close) to the VALVE SEAT width; and second, to blend the valve seat's angle to the valve head's even more shallow angle, so airflow across it at low (and high) lifts is less disturbed cross the Valve Seat.

BALL FUCLRUM: Refers to a simple "ball" principle of rocker arm mounting upon a Stud Mount principle to hold the rocker body to the cylinder head. The ball, which is actually a half sphere, operates directly upon a matching inside radius formed into the bottom of the rocker body to provide the rocker's central location to pivot on during its reciprocating operation. Such a design provides broad strength across more contact area for long life under OEM applications, where this design gets its roots many decades ago, providing it gets continuous flooding of cooling oil, usually from the push-rod. But its "friction bearing" principle was replaced long ago with various rocker arm designs graduating to some form of needle bearing. The irony to this design and its transition to needle bearing rocker arms, which to this day is completely overlooked, is that it has an inherent self correcting, self centering capability because it can "rock" on two axis, where the needle bearing rocker arm is constrained to only one axis of the rigid tolerances of the close fitting needle bearings. This dual axis freedom allows the Ball Fulcrum rocker body to lean to the side when needed, from the misuse (and incorrectly designed) offset rocker studs of In-Line Valve array cylinder heads, like the Small Block Chevy and Ford, which have increased port widths on most aftermarket designs, with corresponding offset push-rod locations. When the rocker stud gets moved off center from the valve to accommodate the offset pushrod, needle bearing rocker arms [unfortunately] get twisted (see: Twisted Rockers), and the rocker's roller tip does NOT sit squarely across the valve tip. The Ball Fulcrum rocker automatically senses this "compound angle" that has now been induced to the head design, and shifts its pivotal axis accordingly, even swiping diagonally across the new motion plane as needed throughout the valve lift. MEI recognized these virtues in designing its OE Series and PF Series chromoly steel rocker arms, and provided the necessary slot width to facilitate this side axial motion on such offset cylinder head installations, as well as provide using two different stud size ball fulcrums, without changing the rocker body.

BALANCE TUBE: (Exhaust Systems) This term refers to the practice of connecting both sides of an engine's exhaust coming from each head through a common tube that crosses from one bank of exhaust to the other. Also used on engines having only two cylinders, such as motorcycles and other smaller powered vehicles. The main principle behind doing this is to capture negative pressure waves coming from one side to enhance a reduced pressure for succeeding waves that follow from the other side by more or less making them act as one. The fact that on multi-cylinder engines having two banks of cylinders, it is impossible (nearly) to design the system so that each cylinder is in the same sequence of timing ahead or behind another in how the exhaust is routed; often with two adjacent cylinders firing directly behind each other, instead of a more preferred and delayed spacing that will induce these negative pressures at a more preferential time. But the balance tube has proven to make the infrequencies of this cylinder timing less important, and in most cases enhanced improved power and flattened out the torque curve without hampering upper RPMs, as is usually the case with other tuning tricks. The diameter and position down from the head is important. (See: COLLECTOR)

BOLT STRETCH: Is the very specific and precise method of measuring the true value of a fastener's specific resistance in creating a locking force. It is a TENSILE measurement resultant of actually stretching a bolt to a predetermined length from its static (non-load) dimension, through the torque used to secure it.  ^

BOSS 429: Trade name of Ford Motor Company.  Okay, it doesn't really have any place in an "educational" definition of terms in this section. But this engine was the catalyst for many things in racing, including the stepping stones to this company. It presented, arguably, the most revolutionary innovations of cylinder head and valve train design for its period, way back in 1968 when it laid only on a draft table. It also created the most complex valve train issues, which are at the root of MID-LIFT technology's evolution, and has served as the model of precedents still used to this day for most rocker arm design solutions by Miller Engineering Inc.

BRAKE SPECIFIC FUEL CONSUMPTION (BSFC): Refers to the efficiency of fuel consumption, per pound, per horsepower produced. The values need to include WEIGHT and TIME per Horsepower. It is acknowledged that 0.42 lbs//Hr. is a standard reference point per Horsepower. At 6.0 lbs per gallon of fuel, a given engine will burn 0.07 GALLONS of fuel per horsepower, per hour. At 400 Horsepower, that would be: 400 x 0.42= 168 lbs / 6 lbs (per gallon)= 28 Gallons/hr. Air density, humidity and temperature outside of standard conditions will vary this mean value.  ^

BYPASS VALVE: (Turbocharger) Sometimes referred to as a "Blow-Off Valve" (BOV), this is a check valve with a predetermined pressure limit designed to release excessive amounts of "boost" from the Turbocharger's Compressor feeding the engine, to prevent over-boost and engine damage.

CAM: An eccentric shaped circle that has a base diameter which is less than 360 degrees that rotates about an axis which increases its diameter from the axis as it rotates through a 360 degree rotation. Generally egg shaped on designs that work with a direct contact running surface "cam follower," or having more squared radius` for designs where the cam follower incorporates a roller ended running surface. See: CAM FOLLOWER.

CAM BASE CIRCLE: (a.k.a. The "HEEL") The lowest point of measurement to the axis of rotation, this is the constant radius, NON-eccentric running surface of the camshaft that has NO effect on lift. It is also the mathematical foundation, as measured in degrees of rotation, for the starting and stopping points of where the lash ramps and acceleration ramps begin and end. However, these specific measuring points are gauged from a relative point, marked as a value of lift from the instant tangent point of where the ramps meet the true base circle. These points have been different over various cam companies and years of cam evolution, where tappet lift measurements such as .010", .012", .015" , 020" and so on were used to describe "gross duration." But the "standard" by which most real comparisons between cams is made, is marked at .050" tappet rise off "base circle." This higher tappet lift measuring point more equally compares two cams without the influence of lash ramps, which often can stretch out to look like a big cam, when measuring in this gross duration way, but in fact may be very tame by comparison to their .050" figures.  ^

CAM CARD: The printed record of cam specifications provided with a camshaft to give fundamental design information necessary for selection, use and installation within an engine. This information should include CAM LIFT, CAM DURATION (often shown in two values, but most important is the .050" Tappet Lift), CAM LOBE SEPARATION, CAM PHASING (installed timing with crankshaft) and VALVE LASH (if mechanical). Additional information often included is "Theoretical" VALVE LIFT (based on specific rocker ratio, often not accurate to reality) and "Theoretical" DURATION (often taken at irrelevant lifter values close to the closed valve point, such as .012", .015" or .020" Tappet Lift; and usually of most value to hydraulic grinds, but often not considered of much value to experienced engine builders).

CAM CENTERLINE: Actually has two meanings. The usual meaning, refers to the cam's designed INTAKE centerline position, in CRANK degrees, as measured with the CRANKSHAFT position AFTER TOP DEAD CENTER (ATDC) of the piston. It is also known as the specification which cam "PHASING" is used. Again, the advancing or retarding of the camshaft with the crank's timing, as per the cam manufacturer's design specifications. (See: CAM LOBE SEPARATION) The second meaning, pertains to the more obvious implications of its name, meaning the actual centerline running through the base circle's axis, up through the cam's nose, thus dividing the opening and closing events equally. Even this though, can be misleading, since many asymmetrical cams don't have nose specifications that are evenly divided. None the less, there is a theoretical centerline to the computer's data which calculated and assembled the CAM RAMP dynamics.

CAM DURATION: This the period of "time" in which the CAM LOBE is active beneath the CAM FOLLOWER, as measured in one of two ways: degrees of rotation at the cam, or degrees of rotation at the crankshaft, the more preferred standard, with regard to engine builder terminology. Cam Duration historically has been measured from various, non-conforming reference points of tappet lift for its "gross duration," which technically is the total duration measured from the instant tangent point of cam ramp departure from, and returning back to the BASE CIRCLE. Gross duration has also been referred to on cam cards in different ways by various cam manufacturers who choose various starting (and ending) points for where this measurement is taken, always termed in thousandths of an inch rise from base circle; such as .012", .015", .020" or more of tappet lift on both opening and closing sides of the cam. The more this figure becomes, the less "gross" it is. So these higher measurements were more often used for timing purposes to install the cam with the crank by, or to compare one design to another for the effective duration (actually seen by the cylinder heads). But thankfully this comparison reference point has settled down in the last 20 years or so to including the more preferred .050" TAPPET LIFT reference. From .050" (fifty-thousandths of an inch) LIFT from BASE CIRCLE on BOTH Opening and Closing sides of the cam lobe, CRANK degree measurements are used to mark a particular cam's most active profile.

CAM FOLLOWER: (TAPPET or LIFTER) A cylinder shaped component that rides upon the cam lobe of an OHV engine, and accepts a PUSHROD on the opposite end which provides the reaching link to the rocker arm atop the cylinder head. Cam followers are either hydraulic "roller," hydraulic "mechanical," or mechanical roller or straight mechanical, aka "SOLIDS." The ultimate performing cam follower is a SOLID ROLLER (no hydraulic valve lash preload); often simply referred to as a "roller." A roller lifter uses a hardened roller race which runs directly upon the cam lobe in conjunction with hardened needle rollers to create a true needle bearing of high strength and impact resistance. It is permanently mounted within the tappet's housing with an interference fit solid pin, that may or may not have additional locking retaining rings set in grooves for added security. The ROLLER Tappet requires a dedicated CAM lobe, for both the hardened surfaces required for mating to the hardened roller's race; but most specifically in corresponding with the angular acceleration formulas which a cam ramp mating to a constantly rotating tangent point requires to keep the acceleration dynamics of the tappet "in-line" with the acceleration rates pushing up from the cam centerline. The simple way to understand this, is that the pressure point of a cam lobe contacting the roller is shifting around the roller as the roller follows the cam lobe up and around the top (cam nose). Whereby a SOLID tappet which has a "fixed" surface laying directly atop the cam lobe is always receiving its upward lift in a straight line parallel with the tappet body, away from the cam's centerline.  ^

NOTE: The velocity of the flat tappet lifter/cam is limited to a specific value by the DIAMETER of the tappet body and the BASE CIRCLE of the cam in conjunction with the LOBE LIFT being used. Since the rubbing surface on the TAPPET'S bottom side receives its "lift" by the nose of the cam sweeping across most of its entire face (contact surface) in the process of rotating beneath it, if the LOBE LIFT is too great or has too much angle coming off the BASE CIRCLE in its attempt to create more velocity, then the LOBE literally sweeps outside the diameter of the tappet. Accordingly, the tappet is no longer running atop the cam if this has happened, and destruction occurs.

The ROLLER tappet does not have this problem because the point of contact between the cam lobe and the roller is always moving around the axis of the roller tappets roller centerline as it follows up and down with each rotation of the cam lobe. This concept allows nearly no limits beyond those established by the laws of inertia, mass (weight) acceleration and harmonics reaching their limits first.

MUSHROOM TAPPETS (slang) are a variation of the traditional, one diameter body dimension that rides within the block's tappet bore, with the exception that the bottom "face" diameter was increased to a larger, "stepped" diameter that gave rise to the term "mushroom." This is a benefit only where the increased cam velocity and/or duration is needed for a particular engine's cylinder head, rpm and cubic inch combination; and where the ROLLER tappets are prohibited by class rules. Mushroom tappets were developed so their larger contact diameter would allow more aggressive cam rates (faster) for the base circles being used, which are always limited to the engine's particular cam bearing diameter. In point of fact, the mushroom lifter goes back many decades, where it was used on older and simpler engine designs (not requiring high cam rates), that simply wanted the lighter weight, and/or a smaller diameter tappet body, while still incorporating a larger contact face diameter to function as needed on the cam lobe. Racing applied the concept to the conventional design many years later as a loop hole technique to get around "non-roller" tappet cam rules for racing which required that only a flat tappet cam be used, particularly STOCK CAR racing (as it used to be called). Today, some classes of racing allow the mushroom tappet, and some don't. On the more popular Chevrolet and Ford engines where such detours in class rules allow, but prohibit the Mushroom tappet, the Chevrolet tappet bore of .843" is often bored out for the larger .875" diameter Ford tappet to accomplish the higher cam rate goals. This mere 1/32" can make a world of difference to increasing the cam rate.

CAM HEEL: See: CAM BASE CIRCLE.

CAM LIFT: This the actual measurement of the cam lobe's total rise from the BASE CIRCLE of the cam which is imposed on the CAM FOLLOWER. Cam lift on OHV engines is always a lower value of net VALVE LIFT because ROCKER RATIO multiplies this by 1.5:1 or more in most cases.

CAM LOBE: The individual eccentric of a cam SHAFT that drives each valve through other components or directly.  ^

CAM LOBE SEPARATION: Pertains to the physical separation of cam lobe centerlines, in CAM degrees. This is also the reverse perspective of "OVERLAP." Both terms deal with the physical timing between the intake and exhaust lobes of a camshaft for the given cylinder in question. Traditionally, OVERLAP was the term used to deal with cam tuning to cylinder heads and cubic inch needs on a given cylinder head. In certain circles this still is. But LOBE SEPARATION became prevalent in the late 1970's as lower compression ratios became the norm from the Detroit manufacturers, adapting to more stringent EPA regulations. Increasing LOBE SEPARATION (which decreases overlap) is a way of building higher cylinder pressure on lower compression engines. See Net Effective Cylinder Pressure (NECP).

CAM NOSE: The highest portion of the cam lobe eccentric where valve lift is at its greatest.

CAM OVERLAP: Refers to the time when both, the intake and exhaust valves of a given cylinder are open at the same time, measured in CAM DEGREES. Specifically, it is when the INTAKE valve begins to open, as the EXHAUST valve is just closing. The piston is still rising UP, chasing the exhaust valve closed, and the intake valve is opening to meet it. Increased overlap is needed on higher RPMs to provide additional time for the cylinder to "purge" itself of leftover gases from the prior combustion process. Physically, as explained in LOBE SEPARATION, the overlap cycle is the opposite perspective of the same consequence to cam design. They effect the engine's "running" COMPRESSION RATIO. You are merely saying the same thing from opposite perspectives. OVERLAP is a term not used much anymore, but was very much the perspective of cam development and comparison by professional engine builders during the pivotal cam development era of the late 1960's and 70's, where many of the breakthroughs of cam design were established. The transition to "lobe separation" lingo was the result of marketing by cam companies, beginning in the mid to late 1970's, who needed to boost cam sales and horsepower on LOW compression engines, resultant from the change in Detroit's gas guzzling car production, and ending the Muscle Car era as we knew it. Reduced CAM OVERLAP is INCREASED LOBE SEPARATION. This reduced cam overlap increased NECP (net effective cylinder pressure) within the engine, often boosting factory engine horsepower by in effect increasing the running compression on these milder, low compression engines. The greater OVERLAP was not needed on these low compression engines since they were not going to operate at high RPMs.  Among advanced cam analysis in professional racing engines, cam overlap is still a critical perspective and term to be documented and used, particularly with cylinder head flow analysis.

NOTE: It is important to understand that you need to increase the static compression ratio of pistons (i.e., 9:1 up to 12:1) because a cam of greater overlap will lose "running compression" (NECP). Increasing the piston's compression ratio replaces this lost operating pressure from the increased cam overlap. (Assuming you have cylinder heads capable of breathing at these higher valve lifts and RPMs.) You cannot increase cam overlap and duration on a low compression engine without changing the pistons for higher static compression, and still achieve full potential. Nor can you install larger compression pistons without changing to a cam of higher overlap (and duration), since now you will have too much "running compression" (NECP) and likely burn pistons, valves and create other problems from excessive heat, pre-ignition and detonation. The two run hand in hand. The only exception to these necessities being chosen only for higher RPM operation would be a large cubic inch engine with a large stroke crankshaft; as this increased crank stroke increases PISTON VELOCITY and increases airflow through the heads at lower RPMs.  ^

CAM RAMPS: Opening and Closing, as they are known, are the connecting portions of the cam's running surface that lies between the BASE CIRCLE and the NOSE. Sometimes referred to as the cam's FLANKS. The cam's LASH RAMP is also included here, and are designed to make the transition between the base circle and the more aggressive (quicker velocity rise) "ACCELERATION RAMPS."

CAMSHAFT: The multi-lobe shaft consisting of the required number of eccentric cam lobes needed to operate the valve train, as it is timed by a chain or belt with the crankshaft. It is made of a variety of materials, with 8620 steel being the most common on high performance roller cam engines. Although numerous high tech coatings and other manufacturing methods are changing this norm on certain applications. NASCAR breaks the rules on cost and technology in this regard. CAMSHAFTS of today's high performance OHV (Over-Head-Valve) engine technology are the "heart" of a FIVE (5) component system that makes up the valve-train, which is arguably an antiquated principle of design, dating back nearly a century. Aside from the last component in our list: the "VALVE," these other three components are respectively known as the LIFTER (or tappet, or cam follower), the PUSH-ROD and the ROCKER ARM.

CAM VELOCITY: Is an often used term by cam companies to describe, in a general way, the cam lobe's rate of lift over other profiles. It of course has a very specific engineering value to cam designers, which is given a value of thousandths of an inch per degree of either cam rotation, or crank rotation (double the cam value). "RATE of ACCELERATION" is another general term used for the same meaning, usually pertaining to cams; but we often use the term for valve acceleration.

CANTED VALVE: This term relates to engines which use a 3 plane valve centerline (not to be confused with valve FACE angles of valve jobs). The Chevrolet big block and 351 Cleveland and 460 Ford are just three examples which have their valves leaning on three separate planes, as compared to the cylinder bores of the engine block, for instance. Small block Chevrolets and 302 Fords, 340 Chryslers, etc., are on a "2 plane" valve array; meaning the valve's inclination to the cylinder bores can be described on 2 dimensions - with the 3rd dimension being zero - since they are all in-line with each other when viewed from an end perspective of the cylinder head.  ^

CENTERLINE: Self descriptive, with rocker arms we are referring to the true center of trunnions, shafts, roller pins (and roller), bearing, valve and/or pushrod.

CENTERLINE, CAM: See: CAM CENTERLINE.

CFM: Usually cylinder head lingo; it's an acronym for CUBIC FEET per MINUTE. Most relative to AIR FLOW, measured through a passage, whether it is a cylinder head port, intake manifold, carburetor, fuel injection metering body, or whatever. It requires a specific pressure value for measuring, known as FLOW DEPRESSION. Without knowing the flow depression used to measure anything that has a CFM rating to it, such as cylinder heads, the value is meaningless for comparison. Too often, cylinder heads are promoted with a CFM rating acquired from flow testing, and no flow depression value is mentioned. So comparing to another head is useless and misleading.

CLOSING THE VALVE: Unfortunately, there are well known cam companies that promote an incorrect philosophy of rocker arm geometry; one of using the rocker for a tuning aid to the cam's acceleration rates. Cam acceleration is nowhere near its threshold of velocities which require the rocker arm to be set in a way that the valve's velocity off the seat is artificially accelerated through what is called accelerated ratios. Unfortunately, this is exactly what is promoted by some companies. Granted, there is no single way to do anything; and there may be tests which can show an improvement in horsepower by such rocker arm juggling; but what this really is - is more of an indicator to what the cam design needs to be at a given lift range, rather than an answer to a problem. Using the rocker arm to induce additional valve acceleration through juggled geometry makes properly controlling the test data impossible, regardless of whether the results are good or bad. Additionally, in all cases of doing this, there is LOST horsepower from the excessive motion the rocker arm and valve train are now involved in. (See: OVER-ARCING)  ^

COIL BIND: (Cylinder head lingo) Refers to the collapsed dimension of VALVE SPRINGS. COIL BIND CLEARANCE is a necessary dimension that must be treated with exacting care and attention to cylinder head preparation. This minimum clearance varies with engine builders and some cam companies. Often taking the shallow side of safety. The type of valve springs chosen (single coil, double coil, triple coil) have specific cautions to what this minimum clearance is. Collectively, when all coils are assembled with the actual retainer to be used (because the steps for the inner springs can vary in height), this minimum dimension is acknowledged to be .050". Trust me: do yourself a favor and make it .100". For more specifics, see: VALVE SPRINGS & PRESSURE.

COLLECTOR: (Exhaust Systems) This refers to the larger diameter tube which is used to "collect" the PRIMARY TUBES of an exhaust system's HEADER. Primary tubes usually range in size on V6 and V8 engines from 1.375" to 2.250", or more on radical and large displacement applications. While the "collector" will usually range from 3.00" to 4.50" (or more). The collector's main benefit is to give the high pressure waves exiting each primary tube a common chamber that captures and uses the low (negative) pressure following each wave to enhance the scavenging of succeeding cylinders before these exhaust pulses reach the higher pressure outside (ambient) atmosphere. (See: VOLUMETRIC EFFICIENCY) The collector is often where (or close to where) a BALANCE TUBE is used.

COMBUSTION CHAMBER: (Cylinder Heads) The cast or machined area of a cylinder head which (1) seals the top of the engine block's cylinder, (2) contains the valves (on OHV and OHC engines) and spark plug (or plugs) on non-diesel fuel engines, and (3) most importantly provides a specific shape that compliments the necessary blending of angles to and from the valves, as well as (4) controls efficient flame propagation of the fuel ignition process. These shapes are varied, but follow two basic concepts dictated through the internal combustion engine's history that have been influenced by the type of valve and port angle configurations. These two basic shapes are known as WEDGE and HEMI (short for hemispherical). More advanced designs use more emphasis on cross flow dynamics that occur during the overlap phase of valve timing, where exhaust port flow depressions are used with specifically shaped combustion chambers having various figure "8" appearances follow SWIRL concepts, rather than random or previously ignored flow turbulent flow of simpler designs.  ^

COMPOUND ANGLES: (Cylinder head lingo) Almost self describing, compound angles are the combination of three planes of dimensions (X, Y, and a straddled dimension between them ["Z" is the valve travel direction]. Most prevalent with canted valve cylinder heads, these include big block Chevrolets, 351 Cleveland and 460 Ford engines, among others. Compound angles are frequently used on pushrods, although the corresponding head may only be a 2 plane design; case in point: Chrysler Hemi. Here the valves are dimensioned in 2 planes, but the "offsets" needed for the pushrod clearance around the ports lean into a compound angle (compared to their respective lifter bores). (See: Compound Geometry)

COMPRESSION: Any force applied to or resultant from a collapsing pressure. The opposite of TENSILE.

COMPRESSION RATIO: Is a static (fixed) measurement of cubic centimeter (cc's) volume (or Cubic Inch) displacement between the Bottom Dead Center value of a given cylinder's size (piston down), plus the volume of the combustion chamber and head gasket; in comparison to the compressed volume of the Top Dead Center value (piston up) of the combustion chamber (and head gasket) by themselves. This is ONLY a reference value for building and analyzing engine specifications with. It incorrectly presumes (to the novice) that valve timing to let air into the engine also opens and closes with the perfect top dead center and bottom dead center positions of the pistons. It of course does not. Because of delayed reactions from weight of air, mass, inertia, reflection and other dynamics, the valves open before top dead center and close after bottom dead center. So this theoretical perfect TDC and BDC formula is meaningless in heavy data acquisition of cylinder head efficiency and cam timing. The REAL compression ratio an engine runs at is more like 2/3 of whatever the STATIC definition here is. Known more accurately, as the Net Effective Cylinder Pressure (NECP).  ^

COMPRESSION HEIGHT: (Pistons) Used to describe the height of the piston's top deck from the wrist pin centerline. The piston's "deck" is the main flat TOP of piston that meets the piston's side where the rings are. On some pistons, such as hemispherical, there may usually be a protruding piston dome extending up from this tangent point, leaving no specific "flat" to be referenced. But the corner where this design reference resides is still the same, and compression height is a term all engine builders should know who work with piston manufacturers on custom designs, especially when calculating connecting rod lengths and crankshaft stroke combinations for engine blocks of a known DECK HEIGHT.

COMPRESSOR: (Turbochargers) The cool section housing of a Turbocharger where the exhaust driven Turbine drives the compressor wheel (defined below) within the compressor, which then accelerates compressed ambient intake side air from an inlet side of the cool section through to the outlet side and into the engine's induction system.

COMPRESSOR INLET TEMPERATURE (CIT): (Turbochargers) This is the actual temperature as measured just before the Turbocharger's Cool Section inlet, coming usually from the air cleaner, which should be the same as ambient (normal outside temperature of the moment), unless influenced by a supplement system and/or heat radiated by adjacent components.

COMPRESSOR DISCHARGE TEMPERATURE (CDT): (Turbochargers) This is the actual temperature as measured directly after the Turbocharger's Cool Section outlet that is then fed to the engine's induction system, usually after passing through an Aftercooler designed to bring these temperatures back down by 50% to 60% of difference with the CIT. This increased temperature is the natural result of the Turbocharger's compressing of ambient air to a higher "boost" pressure.

COMPRESSOR WHEEL: (Turbochargers) The carefully designed impeller of the Turbocharger's cool section side, known as the compressor. The compressor wheel has two distinctly separate designed stages to its blade geometry, whereby the inlet side of the compressor housing feeds inlet blades known as the INDUCER, which takes the incoming air, compresses it and directs it through the integral exit blades, known as the EXDUCER.

COOL SECTION: (Turbocharger) This is the Aluminum housing half of the Turbocharger assembly, which contains the Compressor Wheel which accommodates the pressurizing of normal outside air into the higher pressure boosted air fed to the engine. It is called the Cool Section because although it is attached to the other half of the Turbocharger in an integral, but separable way, by comparison to the often glowing red 1,500° plus temperatures of the Hot Section, it is relatively sane, at only a fraction of that.  See: TURBOCHARGER.  ^

CROSS-SECTIONAL-AREA (CSA): Head lingo, that refers to the actual cubic inch (usually) measurement of area for a port in a head, or intake manifold passage (or any other device flowing air, gasses or fluids). Cubic Centimeters (CC's) is the other value of measurement, but rarely used in U.S. development, EXCEPT in the measurement of cylinder head combustion chambers, piston domes and cylinder's inch conversions to CC's with calculating NET static COMPRESSION RATIO. The CSA calculation is NOT derived from a simple width times height formula. It is based from the square root of a value of area. A simple way to get it, which is good for aspiring head developers who slept through all the advanced math classes in school, depends on which way you are going to do the calculation. If you are taking a PORT and trying to find the CSA, you want to turn the shape into a CIRCUMFERENCE dimension; then DIVIDE by FOUR; then MULTIPLY this figure times itself. (i.e., A small block Chevy might have a port that is 1.25" wide, by 2.125" tall. There are TWO sides to EACH of these dimensions. So pretend you are a piece of string wrapping around the inside space of these dimensions and you will have a TOTAL LENGTH (circumference) of 6.75". Now, DIVIDE this by FOUR (4) to get our "square" for a square root multiple, and you have 1.6875". Now multiply this figure times itself (the essence of the square root), and you get: 2.847". That's the Cross-sectional-area. Now what if you are doing the opposite, and taking something that has a ROUND value already, and needing to find out the cross sectional area of it? Lets use the VALVE. Lets say our small block Chevy head has a valve diameter of 2.100". Here you need to use the standard, Pi (3.1416) to change this circumference into our "string" measurement wrapping around it's circumference (not diameter). In this case that would be 2.100" x 3.1416= 6.59736. (To be exact.) Now divide this by four (4), and you get 1.64934. Multiply this times itself (1.64934 X 1.64934) and you get: 2.720". It's easy. All you are doing is converting shapes into a circumference, then dividing that into an even space of four equal parts, then you multiply one of those equal parts times itself. You are creating a four side box, and multiplying the width times the length. But this ONLY works for EQUAL sides of FOUR. That's why if you multiplied the 1.25" X 2.125" of our sample intake port, you'd have come up with 2.656", and this wouldn't have been the actual cross-sectional-area. It was 2.847", as shown above. In case you haven't noticed, I have also given you aspiring head developers a simple, easy way to check out ANY cylinder head, no matter how complicated the shapes. You need to use PORT MOLDS from prosthetic foam, available at hospital suppliers, and such. After you've made a mold of the port, you can wrap a string around various parts of the shape for your circumference measurement. Then do the math above, to compare with the valve sizes. Here's a rule of thumb. The port should be 90% to 100% of the valve's cross sectional area. With the VALVE BOWL area about 10% smaller on HEMI style engines, and 12% - 15% smaller on Wedge engines. The intake manifold should follow a 95% to 100% value of the valve's cross-sectional-area too. This formula is for MAXIMUM EFFORT, high RPM engines. Torque builders would be considerably more moderate. All these variables can be their own book. My purpose here was to answer the question, then give those who really need to know this term a yardstick of how to apply it. Anyone not interested in cylinder head science isn't going to go past the first sentence of this definition anyway. See: TURBULENCE.  ^

NOTE: Cross sectional area is the ultimate precedent for proper evaluation of port flow QUALITY. I've said it a million times, you can give a monkey a hand grinder and turn him loose making a sewer pipe out of a cylinder head port; and the flow bench will have (at some valve lift figure) a ridiculous CFM value. But it will MEAN NOTHING. Because QUALITY AIR FLOW (that always makes horsepower - AND torque) comes from the greatest amount of airflow through the least cross-sectional area port. This is where "VELOCITY" comes from, as well as controlling LAMINAR flow, and preventing TURBULENT flow. Port shapes are not as important as cross-sectional-area; but they are important to controlling the changes in cross-sectional-area. Think about it.

CUBIC CENTIMETERS ("CCs", metric): A value we use in combustion chamber volume, piston dome volume, and deck volume during the process of determining static compression ratio. The CUBIC CENTIMETER to CUBIC INCH conversion formula is: 0.06102; while the CUBIC INCH to CUBIC CENTIMETER formula is: 16.3871. Unfortunately, CC's are also used to describe "port volume" which in the context they are used by many manufacturers and engine builders, is a useless value, as it is usually attached to the promotion or comparison of one head over another as if to correlate to "flow values," and this is incorrect and misleading. See: PORT VOLUME, for more perspective on this subject.

CUBIC INCHES ("CI"): Normally used to determine engine size. The formula for this is as follows: BORE x BORE x STROKE x .7854 x [number of cylinders]. After you calculate through to the .7854 value, you'll have the cubic inches of ONE cylinder; now just choose 2, 4, 6, 8, or however many cylinders the engine is to get the total value for the engine.

CYLINDER HEAD: The essence of what an "internal combustion engine" is all about; the cylinder heads ARE THE ENGINE. This is of course one of the terms added to this section that created the opening note about including "rudimentary terms" to this section, and I hope if you're reading this web site, you've already got a good idea of what a cylinder head is; but I've added this definition mainly to emphasize that all decisions about valve trains, induction systems and even cubic inches, as well as the end performance application for the engine... ALL come from the decisions made about the cylinder heads. Unfortunately, these integral components are often selected before the cylinder heads are chose or developed into their final modified form; with valve lengths too often being one of those choices. You CAN NOT decide Valve Lengths unless (a) you have a prior design to copy and know the results of spring heights, retainer to rocker body clearance and effects of valve tip to rocker roller position; or (b) unless you have the head first and you've checked all these things FIRST before ordering. DO NOT get locked into the sales hype of someone on the phone telling you that "100 long" is what you need. Don't buy any length that is longer than what you really need; and you won't know this unless you've been careful about checking these things just mentioned.  ^

CHOOSING VALVE LIFT: Engine builders who are too often told what the airflow of a given cylinder head is while reciting where the peak flow occurs, and so cam and valve lift is chosen based on this ultimate point where there is no increase in flow for any additional lift of the valve. This is WRONG. What you want to do is the following:

1. Look closely at the flow graph of both INTAKE and EXHAUST ports, as measured in at least 0.050" increments. (0.100" is too big a jump.)

2. On the intake port, calculate and note the percentage of change from each 0.050" increase in valve lift. This will be significant through the middle valve lift range, and start slowing down with less increase as the port nears its peak.

3. When the port "doesn't make sense" anymore, by giving you only 2% or 3% increase in air flow when you increased the valve lift 8%, or 10%, or 15% from the previous measurement, then the port is not being efficient in proportions to the wear and tear you are adding to the valve train. You'll see 15% and 20% and more, jumps between measurements, until the last .100" of "usable" valve lift. Then you will only see 6% or 8%, which is when the port is just about dead, then you'll be down to only 1% or 2% or so. And the port died about 0.050" before then! This is when you want to pull back before you send the valve higher than it needs to go. When the port has already completed 80% of its work, you're only getting 2% gain of the REMAINING 20%, not the entire valve lift range. The stress exerted on the springs and other valve train components for this paltry amount of flow that the engine never can use is not worth it. It doesn't make sense to open the valve another 10%, to only get 2%, when all your gains was double-digit for each step before this point.

Will the head use another 0.050" valve lift to give you 2% more air flow? Yes, but the engine won't. Here's why. When then INTAKE valve is really working, 80% to 90% of its cylinder filling to the "engine" comes from the valve lift between roughly .200" (less valve lift still has the piston rising) and 80% to 90% of its peak valve lift, as defined by when the port efficiency quits working. This point is when you have nearly nothing for percentage of increase in CFM, compared to your measured jumps from one point to another of valve lift, which can be 0.025" or 0.050" per measurement, as soon as the port stops delivering these powerful and "responsive" flow movements in proportion to the valve lift, the return in results is nearly nothing compared to the much higher stress on valve  train components. Remember, that last 0.050" you may have added, to only get 2% more flow, when 90% of the port work has already been done, may very well have pushed you into the red zone on your spring's coil bind clearance, which decreases the RPM limit before the spring goes into the first wave of harmonics. (You always want a surplus of coil bind; more than 0.050". The springs will love you for it.)

On the EXHAUST valve, it's even more important, because 90%+ of its work is done in the first 70% of its lift, beginning (unlike the intake valve) instantly when it opens. You want to use the same formula of making sense. When the last 0.050" increase in valve lift is only giving you 2% or 3% increase in flow, DON'T go to that last .050". Stop the valve lift short of it, and call that "peak flow," because it is!

All of the above is meant for serious competition heads, and less. When you have a "pro stock" level cylinder head on the flow bench, you will see double digit gains way up in the high valve lift range, .800"+ There are formulas between VALVE SIZE and PORT SIZE that are what determine these percentage of gains per each station of lift you check them at. But I wanted the above to be used for a rule of thumb to dispel the incorrect use of too much valve lift on "out of the box" CNC or aftermarket castings which have these advertised flow figures, and wherever the last figure stops, YOU THINK that is where your valve lift should be. IT ISN'T. By the way, don't trust anyone's advertised anything, pay someone you trust to flow the heads for you! Just because a port has been CNC machined, or hand ported, doesn't mean that the cross sectional area from beginning to end matches the valve diameter and valve bowl design for a truly laminar flow port where the flow bench literally "blows you away." Those kind of heads, those really good ones, move serious double digit increases all the way up the ladder, without trading off low lift flow and related velocity going with it. Anyone can take a hand grinder and make a sewer pipe out of heads, but the real test in their quality is the smooth gains in flow from increment to increment. When this flattens out to less than 10% increase from the previous 0.050", the port has reached its limit. STOP. Good luck.

DECK: Usually refers to one of either, the CYLINDER HEAD or the ENGINE BLOCK (cylinder block). It is the flat, gasket sealed mating surface which adjoins the HEAD and BLOCK together; and in many engine designs is a critical datum plane (measuring reference) for corresponding dimensions to either or both.

DECK HEIGHT: (Engine Building) Refers to the actual height of the block's deck from the crankshaft centerline. But the term is also used to distinguish the "piston's" deck height beneath the block's deck when the piston is at TDC.

DEGREE-IN: aka: Cam degree-in, or degree-in the cam. See: PHASING.  ^

DEGREE WHEEL: A circular plate made of any number of materials, being of anywhere from 4" and over in diameter, and precisely marked in equal divisions of 360 degrees along its circumference for providing exact measurement of cam timing when mounted upon the end of the crankshaft with a pointer that is synchronized to perfect TDC of the #1 Piston position. A necessary tool for all serious engine builders.

DETONATION: (Cylinder Heads) The more extreme symptom and next stage of PRE-IGNITION, and also the most destructive form of uncontrolled, nearly spontaneous ignition within the cylinder, which is not initiated by the spark plug. Main destruction is to pistons and rings, where the explosive detonation and cylinder pressures occur against the piston's rising motion. See: FLAME PROPAGATION.

DESIGN GEOMETRY: See: GEOMETRY, DESIGN.

DEGREES of ROTATION: A reference to rocker arms, whereby the rocker arm's reciprocating motion (typically referred to in linear terms of measurement) is referred to in terms of its angle of operating motion. Useful to calculating its critical MID-LIFT point from different perspectives of setting its Installed Geometry.

DURATION: See: CAM DURATION.

DOHC: (Cylinder head, engine and cam lingo) Meaning DUAL OVER-HEAD CAM. See: SOHC.  ^

ENGINE BUILDER PRICES: MEI broke many rules with the technical information we've made public in this web site, and we thought it appropriate to justify the discount information from which everyone makes critical decisions: "prices." MEI is aimed at supporting the Engine Builders, who make a living building competitive engines, which the technical information in this site is also aimed in getting the most from our products. Many end users, educational students and weekend warriors have also found it an invaluable asset to their interests. In 2007, we removed our LIST price which was a real fair market value for the products we offer, because too many other companies use "list" as a bogus way to sell discounts. Our Racer Net is a real "discounted" value for end users, in particular our rapidly growing PRO-STAND series (especially compared against the most expensive versions from the most advertised brands). For those of you who are the customers of our Engine Builders, we hope you will respect they've spent a lot of time and money to justify the modest mark up they make between ENGINE BUILDER and RACER NET. Remember, you get what you pay for!

EXHAUST GAS TEMPERATURE (EGT): (Naturally Aspirated & Turbocharged) This is a critical temperature for analysis of cylinder air/fuel mixtures and tuning, per cylinder on any engine. Depending on the type of engine design, typical EGT temperatures with optimum air/fuel ratios and BSFC will hover around 1,450° Fahrenheit, give or take. Like any value, where the figures are derived are critical in comparing apples-to-apples; most data acquisition systems recommend that fast reaction probes designed for such temperature ranges be mounted uniformly at the same dimension from one cylinder to another, which should be not less than 4", or more than 6". Of course, unique applications may require innovation here. EGT is sometimes used in a less accurate way on less sophisticated engine instrument systems, where only a single probe is used in the final exhaust path, rather than on each cylinder, usually after the turbocharger outlet.

EXDUCER: (Turbochargers) See: COMPRESSOR WHEEL.

FATIGUE: Is a term used to describe a real value to metallurgy and alloy properties, although it is not quantified as such. This is because it is the result of microscopic stress that are relative only to each component design, application and environmental operating characteristics. Specific designs of a product are subjected to cyclical tests, where the operating characteristics desired are exceeded in excessive endurance (time of use), operational stresses imposed, or both. Temperature is also a key component to the accuracy required to quantify the results. In summary: fatigue is the result of stress and time acting upon the molecular properties, where flex and temperature fatally disrupt the alloy's molecular integrity. It is especially susceptible and aggravating where machining flaws and STRESS RISERS can be found.

NOTE: ALUMINUM is one of the best materials to use for a rocker arm, because of dampening properties; but it is also one of the few alloys (if not the only) which will FAIL from fatigue at operating loads which are beneath its ultimate tensile and yield strength values. This is the direct result of fatigue, which is not the case with other metals. Other metals, such as steel, will operate indefinitely, if the operating loads are beneath their ultimate tensile and yield values, all else being equal. It is also worth noting here, that fatigue in aluminum is directly influenced by the alignment of the molecular grain, which is established during the shaping process at the mill; whether that is by forging or extruding. It is a generally accepted fact of metallurgy, that aluminum components which are designed to operate with their molecular grain in line with the load forces, will have up to SIX TIMES more resistance to fatigue. It is also why all Miller Engineering PRO-STAND, PRO-SHAFT and PRO-STUD rocker arms have ALWAYS been machined with their molecular grain running lengthwise to the loads. While the biggest names in valve train continue to use "cross grain" extrusion! (Sorry for the ad.)  ^

FLAME PROPAGATION: (Cylinder Heads) Refers to the character of the burning process of compressed air and fuel in the combustion chamber following ignition by the spark plug (or plugs); but specifically: its progression from the point of ignition, across the piston and within the cylinder head's combustion chamber in a "controlled burn." Although the process is measured in milliseconds, it is a "controlled burn," and not an explosion, as some presume. When the burning process reaches a point of random ignition, where the incoming fuel is ignited by something other than the spark plug, either through other "hot spots" within the combustion chamber, or improper (and/or inadequate) fuel mixture, or some other unplanned form of PRE-IGNITION, the malady soon graduates to an explosion which is then called DETONATION. Once pre-ignition graduates to detonation, it becomes critical and destructive.

FLANK: (Cam Lingo) See: CAM RAMPS.

FLOW DEPRESSION: Head lingo, used in relation to flow bench testing of cylinder heads, intake manifolds and passage ways. Flow depression is most often equated with "vacuum," and although this is usually the type of testing sought, it actually refers to either positive or negative pressure (vacuum) as measured in a relative of "differential pressure" from ambient (the pressure we stand and breath within and around us). In true scientific analysis, where everything must be standardized, tests done on machines which are not in controlled atmosphere (and most aren't - unless you're NASA), use a cross reference chart for all data derived in a test, which transposes the actual environment circumstances to what is known as STANDARD; a specific temperature and humidity value at sea level barometric pressure. Computer assisted data acquisition now does this in lieu of a chart, for the more sophisticated machines. Sorry, I digressed. But it is important to understand when comparing flow data between different tests, and especially different machines and/or operators, that without this cross reference standardizing, the results cannot be truly "apples to apples" in comparing. On high powered, high velocity machines, a MERCURY monometer is the reference of measurement; but in the real world of realistic flow test equipment, a WATER monometer (approximately 1/14th as heavy as Mercury) is used for witnessing the amount of flow differences being measured. This test "pressure" (positive or negative) is maintained for all points of valve lift, at a predetermined standard to operate the flow bench by in which flow data (CFM) is derived. One thing that should be noted here, is that FLOW DEPRESSION can be used in a way to exaggerate flow figures for anything being sold using these figures...like CYLINDER HEADS. All you have to do is TEST at a higher value, like 28" of Water Depression, over a test that is done at 25" and the flow figures will be higher. ALWAYS compare CFM values by the FLOW DEPRESSION, otherwise they are meaningless. It would be like measuring horsepower without knowing what the RPM's were for the test.

NOTE: The above directs the definition of FLOW DEPRESSION as used in "testing," but in engine operation, flow depression is more aptly limited to VACUUM, as that created by the piston and cam timing during normal operation, AND in the synthetic manner created by intake port and exhaust tube "tuning" (in conjunction with the cam and specific RPM ranges) where reversion waves, both super-sonic and sub-sonic create a flow depression behind them that can be synchronized to enhance depressions created in the intake system and across the piston for increased VOLUMETRIC EFFICIENCY. See: RAM TUNING.  ^

FULCRUM: A one piece device used predominantly with the O.E.M. design rocker arms, consisting of a hardened, radius bottom ball or ball and stand combination (then termed a PEDESTAL FULCRUM). Having a hole through its center for a mounting stud that affixes the rocker arm to the head, the FULCRUM provides the rotational friction surface base in which the rocker arm pivots upon. Not desirable on high spring loads and high RPM operation, where its friction characteristics limit such use, as well as rocker arm design - since it is primarily a "stamped steel" rocker component. This design requires excessive amounts of oil flooding to adequately carry the heat of any performance use of such a design away from the fulcrum. Never use oil restrictors to the valve train on performance engines using this design.

FULL COMPLIMENT: Bearing lingo, used in describing NEEDLE BEARING designs which have side-by-side needles (rollers). The other design reference opposite this, is NON-Full Compliment. This latter design incorporates a retaining spacer ring within the shell that acts like a cradle to keep the rollers separated from each other, with spacing that varies from design to design; but usually is about half as many rollers as FULL compliment. FULL COMPLIMENT is the most expensive, and is the higher standard of the two options, because of its inherent higher unit loading capabilities. Rocker arms having non-full compliment have made a compromising decision based on costs, usually not proportionate in real savings compared to the loss of operational integrity.  ^

G-Plane™: A Pending Trademark of MILLER ENGINEERING INC (MEI). Pertaining to the MEI PRO-SHAFT®, it is one or more machined faces of the rocker body which incorporate a specific angle to other fixed angles of the cylinder head in question, so that a precise reference measurement can be taken and used by the engine builder to facilitate accurate placement of the rocker arm's pivot height with the valve tip, otherwise known as the INSTALLED GEOMETRY.

G-TooL™: The Patent Pending MEI geometry tool developed for finding the exact length push-rod requirement to set a rocker arm's Installed Geometry, through three steps and two methods, including a means to confirm the accuracy of the cylinder head's mounting stud. G-TooL INSTRUCTIONS (Printed)  Illustrated G-TooL Instructions  (Web)

GAUGE HEIGHT/POINT: Is an engineering term for a specific point of reference to the design of something, often used for tangents to or from angles that need to be measured from a linear reference point. It is often used for many other things beyond engineering to reference procedures for inspection, manufacturing and assembly; especially where these other categories meet or overlap each other. We use reference to it here in describing the difference in measuring points needed for VALVE LENGTH, in discerning between measurements for designing your own length from the valve seat to the valve tip, versus the manufacturers usual term of overall length (OAL).

GEOMETRY, DESIGN: This is the geometrical parameters of which a specific rocker arm is designed, i.e., the dimensional distances between its connecting components - but specifically the "angles" in which they are placed within the rocker arm body; which are influenced by the operational angles of the pushrod to the valve on the engine in question. This correlation directly contributes to or from the translation of the camshaft's information to the valve. Because the tangent point locations and angle associated with the push-rod side of the rocker arm determine the degrees of motion the rocker arm rotates from whatever the cam lift is, we also refer to the Design Geometry as "push-rod geometry."

GEOMETRY, INSTALLED: This is the relationship of the rocker arm to the valve stem tip, as determined by the engine builder; or more specifically, the operating angle the rocker arm works as measured to the angle of valve. Although the rocker arm is in a "fixed" position during operation, which pivots about an axis that is stationary, the "perspective" of how it operates with the valve is determined by its pivoting relationship to the valve stem tip, or its shaft's height to the valve stem tip. Raising or lower this pivoting relationship directly causes the roller to move across the valve stem tip in varying amounts, known as the WEAR PATTERN, depending on where its "starting" point is at. If the rocker arms shaft sits too low to the valve stem tip, it spends more time rolling across the valve before it is pushing it "directly" down; therefore it is slower getting the valve off the seat. See comments for other consequences. NOTE: See VALVE TIP ADJUSTMENT.  ^

GEOMETRY, TRADITIONAL: The 1/3 RULE. Always based on two things: a SHOE TIP (scuff pad) contact surface for depressing upon the valve, and second: measuring the angles of motion from the contact surface of the shoe's contact pad. Measuring the pivot point's angular relationship with the valve through this contact point when the valve is approximately 1/3 of its total lift is how measuring the rocker arm's length and pivot position has been done since roughly when the Wright brothers went to Kitty Hawk. Incorrectly so, especially for roller tip rocker arms.

GROSS DURATION: See: CAM DURATION (above).

GUIDE PLATES: Referring to both valve train and cylinder heads; these simple devices have been used for many decades to keep rocker arm alignment with the valve by restraining the push-rod's side-to-side motion. This concept of rocker arm alignment has often been simple "slots" cut (or cast) into a cylinder head, rather than the separate components that traditionally mounted beneath the rocker mounting studs, stamped from "plate" steel, and heat treated. They DO REQUIRE HEAT TREATED push-rods to operate without wear. We highly recommend this concept of rocker arm alignment over the more recent, OEM cheap way out of using "self aligning" rocker arms. Avoid using self aligning rocker arms whenever possible, regardless who's brands of stud mounted rocker arms you use. SEE: Self Aligning.

HARMONICS: Like music, harmonics denotes a self description of "frequency" and "vibration." Both are true in engines, because all matter has a natural harmonic frequency in which it resonates, all based from fundamental physics of mass and density, as well as material composition. For instance, aluminum has a different harmonic frequency than steel; and titanium has a different frequency than either of these. What is relative about harmonics in valve train, is the interaction of these operating frequencies between the different components which establishes limits to RPMs, valve spring life and even necessary cam designs. Operational loads normal to a competitive engine compound once the threshold of harmonics is reached, requiring several times more effort to resist these increased loads. More simply stated, in valve springs, increased pressures for high RPMs is NOT added because of RPMs, it's added to replace the lost working pressure that harmonics destroys. Changing to a different frequency alloy, such as going from steel to titanium, elevates the frequency (because that is where titanium is at compared to steel: HIGHER) at which these loads compound, so less valve spring pressure is needed to control valve float with titanium valve springs. SIGNIFICANTLY LESS. Rocker arm geometry directly influences the threshold of any alloy harmonics, because their limits are a direct result of several factors, including "motion" and constrained linear loads held tightly to the linear paths of compression. In other words, keeping the push-rods in-line as much as possible. See: VALVE SPRINGS & PRESSURE & PUSHROD HARMONICS.  ^

HEAD: (Slang) See: CYLINDER HEAD.

HEADER: (Exhaust Systems) This refers to the main manifold system attaching directly to the cylinder head's exhaust ports for the purpose of routing exhaust gases from the engine. But "good" headers are designed so that the leading tubes leaving each cylinder, known as "primary tubes," are of an equal length and specific diameter to accommodate a specific tuning of these high pressure, high temperature gases so that the energy created by their exiting the cylinder can be isolated from following cycles of the same cylinder, as well as used to enhance the scavenging of gases from adjacent cylinders, where each of these tubes come together in what is known as a "COLLECTOR." However, the term "header" is sometimes used with exhaust that have no collector, where only primary tubes exit from the cylinder head individually, as depicted in early designs on aircraft engines, boats, etc. Such designs still exist on supercharged engines, often used in drag racing, but are more often referred to in slang as "zoomies," because of their up-swept and short length that zooms backward with the flow of the wind. In supercharged applications the science of tuned lengths to cancel out REVERSION waves and match naturally aspirated flow depressions for high volumetric efficiency doesn't exist, because the forced air induction of the supercharger pretty much overrides all these laws of ambient physics.

HEEL: (Cam Lingo) See: CAM BASE CIRCLE.

HORSEPOWER: In its purest definition is: 1 HP= 33,000 Foot-Pounds of WORK Per Minute.  In measurement, it is TORQUE multiplied by RPM divided by 5252. The essence of how horsepower is derived is based upon an understanding of the difference between WORK and TORQUE. FORCE is the common denominator of the two. WORK is DIRECT FORCE operating across a distance; while TORQUE is RADIAL FORCE operating around an axis. The formula of these is: POWER = FORCE x DISTANCE ÷ TIME. To calculate power, both DISTANCE and TIME must be derived, using FORCE across both of these. Where engines are involved, the rotational values of the crankshaft by its diameter is needed. To generate a formula for the TORQUE, the DISTANCE aspect is measured in RPM (revolutions per minute), times the RADIUS, times Pi (3.1416), times 2 (or 6.2832.). TORQUE is measured as Pounds of Force Times Distance. POWER is measured as FORCE times Distance per Minute. The 5252 value in our equation comes from 33,000 divided by 2x Pi (6.2832). Therefore: HP = Torque x RPM ÷ 5252. Something worth noting, is that at 5252 RPM, both TORQUE and HORSEPOWER will be equal. Beneath this value, Torque will be numerically greater, while above it Horsepower will be greater. Just buy a Dyno!  ^

HORSEPOWER CLAIMS - Rocker Arms: Every rocker arm manufacturer loves to make claims, like "30 horsepower" from switching to their rocker arms. It doesn't matter which rocker or cam company you name, they've all made these claims at some point. We have NEVER made a horsepower claim when asked; even though it is very rare that significant power isn't made with changing to MID-LIFT rockers. But the ads run by other companies speak as if their products by themselves were responsible for whatever power was found, even if power was truly found. But it is not accurate, or we should say: it's not "apples to apples." THE PROBLEM with this, is that there has never been any neutralizing, comparative data to show that the tests were truly done in a way that isolates the substitution of their rocker arm from the previous example rocker, whatever that may be. In other words, there is never any data that accounts for the many other factors which influence what the engine sees in creating this horsepower, as a result of the rocker arm change. What difference does this make, you ask? Well, the ads present an improvement in performance, measured in horsepower, from a mere swap of rocker arms. But what is missing is that the change from a stock rocker to a roller tip rocker by itself, has a different lift dynamic at the valve, than the preceding rocker being compared to. In other words, the "valves" see a different valve lift RATE, probably duration, and maybe LIFT. In other words, the engine had a cam change, as far as the valve lift was concerned. This is where the credit for the horsepower should be placed. Not the rocker arm. So when you read these ads about "30 horsepower" improvement, remember that unless there is additional information stated that makes it clear a comparative test was done where the VALVE LIFT DYNAMICS was duplicated at the valve, then no credit can be fairly given to the rocker arm's features (roller tip, needle bearings, stainless steel, etc.), as the ads most probably imply. Remember, the rocker arm is DOWNSTREAM of the cam profiles, so whatever you do with rocker changes or geometry, directly affects what the valves see! That's what you're measuring. Now you might ask "so what?" You might say "I don't care how my rockers made more power, they still made more power!" But my response would be, that the ads are comparing THEIR rocker arms to OTHER rocker arms, and the OTHER rocker arms MIGHT have given you MORE power too...over the advertised rocker arms, if all else was equal. But it never is equal. That's the point. No one has ever used a "STANDARD" to cover all these variables. The point is, you have to be fair and careful of what you are buying, and why you are buying it. If you want to put more valve lift in your engine without a cam change, then changing rockers with higher ratios is a good way to do it, PROVIDING there is enough piston to valve clearance, and providing you don't lose running compression (from increased overlap) as a result, which then loses horsepower. You need to be clear in understanding the difference in TEST data that claims one thing, but which is the result of something else. That's all.

HOT SECTION: (Turbocharger) This is the heavy, cast iron side of the Turbocharger, which is mounted directly to or integral with the engine's exhaust system, often at the common collector for the engine's primary exhaust tubes. The Hot Section is carefully designed in its internal round shape and cross sectional area to feed the hot, high pressure, high temperature gases from the engine into the TURBINE WHEEL, at high speed, often from 40,000 to around 200,000 RPMs, which drives the COMPRESSOR WHEEL through a common shaft. It is not uncommon for the Turbocharger's Hot Section to get cherry red in many installations, and care to location and insulation (when possible) is needed in controlling this very high radiated heat. See: and TURBOCHARGER.

INCLINED STUD: Most engines have a stud that is inclined to the valve's centerline. With the most popular engines, this dimension varies around the 10 or 12 degree range. That is, the rocker arm stud leans into the valve this amount. The engineers arrive at these figures through several considerations revolving around each engine's particular needs - but the over-riding consideration is to average (or split) the difference of the inclination of the pushrod into the valve. With the small block Chevrolet, the stud leans 11º 20' (11-1/3 degrees). The more you raise the valve stem tip, the higher the rocker arm must be; and the higher you raise the rocker arm - the closer the stud's centerline (and the rocker's pivot length) comes to the valve; making the juggling act of "centering" the roller to the valve nearly impossible from one engine to another. This is one major place (and theory) that most engine builders and manufacturers get into trouble. The logic of "centering" the roller on the valve is indicated as a good method for setting correct geometry, when obviously it's not.  ^

INDUCER: (Turbochargers) Pertains to the COMPRESSOR WHEEL of the Turbocharger's Cool Section, but specifically it is the smaller diameter, and the first stage of two for the complex blade geometry of the impeller. This is the specific part of the compressor wheel which takes in outside air (usually from the air cleaner and connecting tube). After compressing with the corresponding inlet shape of the surrounding Turbocharger's cool side housing, it then forces it outward across larger diameter, integral blades of the same component which are known as the EXDUCER, whereby the air is forced into higher pressure and corresponding temperature (not a virtue) into the engine's induction system, usually after passing through an AFTERCOOLER.

INSTALLED GEOMETRY: See: GEOMETRY, INSTALLED.

KEEPERS: (Cylinder Heads) See: VALVE SPRING RETAINER.

LAMINAR FLOW: (Cylinder Heads) Refers to the quality of airflow through a passage or over a surface. In cylinder heads, it is the quality of airflow through the ports, specifically the INTAKE port. Specifically, Laminar flow is LAYERS of smooth or spiraling airflow that is controlled and consistent. The opposite of this, is TURBULENT airflow, whereby air is reflecting and tumbling in all directions within the port passage. Turbulent airflow on a flowbench can be heard loudly over laminar.

NOTE: Experienced flowbench operators who chart the airflow curve will notice one of two phenomena in the patterns of low valve lift flow up through the point where turbulent airflow overtakes this steady rising increase, and the port flow drops like a rock. The first will be a steady rise in airflow, per step of valve lift chosen, followed by a gradually flattening out of this curve near the port's peak flow limit. When the valve lift has exceeded the port's ability to feed it, the flow will drop off quickly. The second example is a faster rising airflow, per valve lift, followed by a quick (without warning) drop, usually missing this leveling off of the flow curve near the top of its flow capabilities. This second port is the perfect example of a "turbulent" port; often too big for the valve it is feeding, or the valve is too small. When trying to apply the MEF (below), the POINT where a flow value should be taken is about 90% of the valve lift where flow increase STOPS. NOT turbulent, but where the curve is more or less flattened out. If that valve lift point is .750", and full turbulence is .810" (for example), then 90% of the .750" is what we've always chosen; which would be .675", as the CFM value used in the MEF calculations.

LASH: See: VALVE LASH.

LASH RAMP: (Camshaft) Is the area of the acceleration ramp where the cam lobe actually begins lifting the TAPPET from the BASE CIRCLE (or Heel) of the CAM. Although the lash ramp isn't really its own entity of the cam, merely the bridging area of the cam that has the slowest rate of acceleration to allow the increasing rate of tappet lift (in both height and speed) to continue to increase without too much shock that launches the tappet's contact away from the cam's surface. As noted below in LOBBING the VALVE, this was accidentally done in the late-1970's when a well known drag racer found more horsepower by excessive lash approaching .045". This was later pursued by other racing engineers with purposely radical acceleration ramps that were intended to "lob the valve." Both practices: excessive lash, and lobbing the valve with radical ramp speeds, are rarely done or accepted by professional engine builders, although a couple of "rogues" may still lay loose out there. But the most important aspect of both, was to change the limits of what controlled high velocity cam acceleration could be, from what was perceived prior to this turn of accident and events around 1979. The previous milestone of high performance cam breakthrough being when Chrysler contracted Crane around 1971 to develop a new cam for their Chrysler Hemi powered Pro Stock drag racing teams, and the .700" Valve Lift barrier had been broken.

LIFTER: Also "tappet." See: CAM FOLLOWER.

LINEAR MOTION: Meaning "in-line," linear is really self descriptive. Here, it pertains to the devices which are designed to follow an in-line motion (but often don't, because of improper rocker arm design and installed geometry). These two devices are of course the "valve" and the "pushrod," which work through the rocker's RADIAL MOTION.  ^

LOBE SEPARATION: See: CAM LOBE SEPARATION.

LOBBING the VALVE: (Slang, Cam lingo) This is a technique of both cam design and application by the engine builder whereby a specific cam grind known to have acceleration ramps purposely designed with an overly aggressive velocity that will accelerate the cam follower and VALVE over the nose of the cam, literally throwing it into space at a given, terminal velocity from a predictable, high RPM, approaching the threshold of uncontrolled valve float. This can be said to be "controlled valve float." Its origins, like many evolutionary points in racing, was by accident, going back to Drag racing around 1980. It happened from Pro Stock racers finding horsepower by (accidentally) running increased valve lash; approaching .040" to .050" - a previously considered dimension that made valve train destruction imminent. Which it is. But if used for very minimal periods of time (like drag racing), it can be tolerated - with risk. The data from this led cam "thinkers" to design ramps that would approach these previously taboo acceleration rates, but with more conventional valve lash. Rocker geometry is also pursued by some engine builders to induce these dynamics, but again, as stated elsewhere, there are tradeoffs for these second level methods of "tuning" which are unnecessary, considering roller tappet cam designs can still create whatever velocity is required (flat tappet cams have more constrictions - based upon tappet diameter), and the smoother, less damaging motions of the upper valve train can be kept in check. Valve lash settings and lobbing the valve during specific controlled testing, should only be used as a barometer to indicate more specific cam grinds for the heads and performance application.

LONG SIDE RADIUS: Cylinder head lingo, that refers to the top side of the intake or exhaust port, where the bend to (or from) the valve pocket area is made to connect the main port length's passage way, or port "roof."

LONG SLOT ROCKERS: Long slot rockers is a term used to define a modified stamped steel rocker arm body where the adjusting screw slot has been increased to accommodate a longer arcing motion of the rocker arm for more valve lift. The irony here, is why the manufacturers had to make them "long" from where they were, because they were long enough to accommodate the valve lifts being sought. But the real problem is that the slot was IN THE WRONG PLACE to begin with! We use a more appropriate term to describe them: "Wrong Slot Rockers."   ^

MID-LIFT: A Registered ® Trademark, MID-LIFT (HALF-LIFT) pertains to the middle of VALVE LIFT, which is the most critical dimension that must be known to install the rocker arm geometry correctly. But we use this also for the camshaft, which is critical in understanding and timing with setting up precise rocker geometry. You must remember that the rocker arm is a "messenger" of the camshaft, but more importantly, it is a SYMMETRICAL messenger. Which means that it duplicates its motion identically, on both sides of its "rocking". No matter what the geometry of the rocker arm is, whether it is sitting too high and under-arcing, or sitting too low and over-arcing; when it comes time to close, the pattern, the acceleration rate and the angular acceleration will be the exact reverse of whatever it did to open, except as influenced by an asymmetrical cam flank, which is of minor consideration to the geometry point being made here. It is for this reason, that these two motions must be divided equally.

CONSEQUENCES: If MID-LIFT IS NOT utilized, the rocker arm becomes a second variable to the cam's design, directly effecting acceleration rates of the valve. The term "area under the curve" is often used, but less understood. The "curve" in question, is the "lift curve" of the valve. And there are applications when opening the valve off its seat as quick as is possible may not benefit performance. But let's assume it does. Let's assume you want the valve to jump off the seat as fast as possible, to give you that "area-under-the-curve" (big breathing at low valve lifts); then, you slow the valve down as it approaches full lift. All of the above is classic horsepower building "trick" cam design. But if you use the "rocker arm" to make the valve see this, there is a symmetrical reaction to the "closing" cycle of the valve and rocker arm, which SLAMS the valve down with the same velocity that it accelerated away from the seat. If the cam has been designed with the softer closing ramps needed to keep valve springs and valves alive, such rocker geometry has just defeated the dynamics of the cam's design.

Just for the record: Rocker Design and Installation as MID-LIFT dictates was never being done before; not in the U.S., not in Europe; not 30 years ago, not 80 years ago. The "1/3 Rule" explains what was being done before. References to foreign engineers predating the US Auto industry did not address BOTH sides of the rocker arm as the Miller MID-LIFT Patent does. It's a big world, and many great engineering concepts most people think today are "new" were first tried long ago by many automotive (and aircraft) geniuses who've long since passed. MID-LIFT, however, was never one of those "done before" ideas. Prior to the MID-LIFT Patent's issue in 1982, there was never any published, or printed, or explained illustrations of how to measure or setup MID-LIFT rocker arm geometry by anyone, not even on just the valve side (let alone the pushrod side). All information of the very little that occasionally came out was hit and miss generalities that were wrong; many of which still continue to be thrown around today.

MID-LIFT (Wikipedia: Link to the Internet's Dictionary).

MID-LIFT PATENT: A U.S. Patent, covering the concepts of balanced geometry, issued to James M. Miller. It is the only such Patent ever issued for a concept of precision geometry for any rocker arm.  ^

MILLER RACE ENGINEERING (MRE): Original and first shop of Jim Miller; opened in 1977 and specializing in cylinder head and intake manifold development -- originally for the BOSS 429, but quickly grew into high tech unlimited R&D on small block and big block Chevrolets, Hemis, along with Jim's personal preference with the Fords. Drawing work from all over the U.S., thanks in part to Jim's early writing talents for numerous automotive magazines that covered both in-house programs and the latest tips from competing shops alike, Jim was one of SuperFlow's earliest customers in 1975, working closely with Jack Roush, Holman-Moody and many other notable teams. Jim turned down an offer to head up as Crew Chief from Ford Pro Stock Legend, Dyno Don Nicholson in 1975 to open his own shop shortly later.

MINUTES (Angular Terms), See: ANGLES.

MOHAWK: A North American Indian having a unique hair style.

MOMENT of INERTIA: Is the center of gravity equivalent to any mass which, in the case of rocker arms, refers to the starting and stopping effects of this mass in a reciprocating manner about an axis of rotation. The term is unfortunately misused in rocker arm promotion, as though it's being "lower" is some accurate accomplishment of rocker body design. The "lower" moment of inertia which everyone so loosely throws around, actually refers to placing it closer to the center of axis of rotation. It is actually a "symptom" of design by our standards, because, unlike some educational entities, we don't design around it, we try to minimize it with the ultimate least amount of metal required to effectively control the unit loads and fatigue of a rocker body. No one who understood rocker arm design, would purposely add weight to a rocker body, except to correct some design flaw in strength. Moment of Inertia would "consequently" increase with this. But if that's what it took, then so be it. It is NOT a criteria for design, if all other aspect of body design follow minimum material for maximum strength needed.

MOTION LINE: Is the imaginary line to be made, drawing through the rocker arm's shaft or trunnion axis, and the axis of the roller pin. It is this line which is placed at a 90º angle to the valve centerline, at MID-LIFT.

MOTION TRIANGLE: Is a term we use based on the tangents located on the rocker arm, at the centerlines of either the pushrod side or the valve side which creates the THREE POINTS of critical geometry; TWO are from a connection between FULLY CLOSED to FULLY OPEN (cam or valve) and the THIRD is the AXIS POINT of rotation of the rocker arm. (See: MID-LIFT ARCING)

MUSHROOM TAPPET, See: CAM FOLLOWER.

NATURALLY ASPIRATED: (a.k.a. NORMALLY ASPIRATED) Meaning an engine which operates through conventional and traditional breathing created by the sequential cycles of vacuum induced by the pistons in timing with the valves, that allow ambient (natural) air pressure to rush in for compression and ignition. Non-assisted by supercharging, turbo-charging, Nitrous Oxide or any other external means of pressurizing of the cylinders.  ^

NECP: NET EFFECTIVE CYLINDER PRESSURE: This is the REAL running compression ratio of an engine. It is the by-product of combined STATIC compression ratio, and cam/valve timing. As explained in COMPRESSION RATIO, the valve timing in reality, opens and closes beyond the perfect top and bottom stroke positions of the crankshaft, therefore not utilizing the full compression of the static compression value; usually about 2/3  or less (i.e., 12:1 CR = 8:1 NECP).

NET VALVE LIFT: See VALVE LIFT.

NOSE: (Cam Lingo) See: CAM NOSE.

OE SERIES: Refers to the MEI designed "Original Equipment" style, 4340 chromoly steel, "shoe tip" Ball Fulcrum rocker arm, with numerous innovations not found on any other such rocker arm (sorry for the commercial). Developed in 1998 for Precision Valve Systems, Inc., following the introduction of the PF Series rocker arm, the OE Series was aimed at stock rebuild market, but soon became THE "stock type" rocker arm to have for serious Late Model and Stock class competitors who were allowed radical camshafts, but constrained to the principle of stock type rocker arms. The first feature that made this rocker a must have, was the 4340 Chromoly steel which no other shoe tip rocker had, because it was too costly on tooling to make. But another revolutionary concept of the design, broke all engineering rules in adopting the MID-LIFT roller tip rocker principles to the rocker body's tangent points of motion. Simply stated, the rocker arm had a shoe (pad) tip, but acted like it was a roller tip, especially when the instructions were followed to the letter. An easy task, as this rocker arm also had a precisely ground flat top surface, which facilitated a reference plane for measuring with the rocker stud, and allowed use with MEI's Patent Pending G-TooL™.. The result, like any MEI MID-LIFT geometry rocker arm, is that the area-under-the-curve was at its greatest, giving the most valve lift, velocity and duration for each degree of crank motion, throughout the entire valve lift cycle; while showing the same, minimal back-and-forth sweep atop the valve tip. The in-and-out pushrod motion was also nearly non-existent. Made in only one batch of 40,000 pieces for the Chevy SB head, they are used by Ford engine builders too, they were sold out quickly, never made again, and have become like gold, as they are the only such design that lives with professional level cam velocities in class competition which demands only "stock type" rockers be used. OE Series INSTRUCTIONS

OAL: "Over-All Length" refers (as it sounds) to measurements of linear objects, taken usually from their outside dimensions of one end to the other.

OEM: "Original Equipment Manufacturer" refers generally to the original automotive manufacturer, such as Chevrolet, Chrysler, Ford and so on.

OFFSET: This term is usually used to describe the difference of the pushrod alignment with the valve, as the head and rocker arm are looked at from an "over the fender" perspective (head's exhaust ports facing you). As intake ports are widened, there needs to be an offset of the pushrod away from the original port location with some head designs, to provide clearance. An offset guide-plate is the usual method with stud design rocker arms. To this day, it is still common practice to use stud mounted rocker arms which are "twisted" on the stud's axis. TABOO! (See: Twisted Rockers) When this is done, especially on the small block Chevrolets and Fords which have 2 plane valve layouts, you are introducing a 3rd plane angle that will, at some point throughout the lift cycle, LIFT the roller off to one side of the valve tip's edge. (See: COMPOUND GEOMETRY) The roller may lay flat in the closed valve position, but it will definitely leave this alignment during the valve lift cycle. You can see this on an engine that has operated as such, by a ring wear around one edge of the roller (indented ring in worse cases), and the rounding of the valve stem tip itself.  ^

OIL RESTRICTORS: Is the technique of limiting the volume of oil to the top end of the engine through one of several methods, depending on the engine. One method restricts the oil feeding the tappet's oil galleys with plugs in the back of the block, another method simply uses restricted oil hole sizes in the pushrods, but does nothing for limiting a surplus of oil bleeding off from the tappets, if they don't need it. Some restrict in the rocker arm itself, with orifices down to .032" to .043" in diameter. The thinking behind this is to reduce a surplus of oil that is not needed to the upper valve train, so that more is kept in the oil pan (or sump). A key method for doing this is in the main galleys which feed oil to the tappet bores. When Hydraulic tappets have been substituted for mechanical, the need for oil pressure is not as mandatory to keep the hydraulic valve within the tappet in a compressed state. This is reason number one. Reason number two, is the belief that excessive amounts of oil going to the top of an engine which is only used for short periods of time (such as drag racing), is a waste of resources from the crank's rod and main bearings. Reason number three, is "I got needle bearing rocker's and I heard I don't need that extra oil up there." In most cases, all three are wrong.

NOTE: With the PVS OE Series rockers and their friction "ball" fulcrum, this is a warranty killer; and totally taboo. The factory oil systems are fine on 99% of all the small block Chevrolet and Ford engines (big block too). The only shortcoming "might" be the gear-to-gear oil pump which Chevrolet uses. Ford's gyrator design is undisputed in self priming, high volume at low RPMs and stable flow at high RPMs. But the history of "oil restrictors" dates back to  Bill Grumpy Jenkins Pro Stock small block of the late 1960's and early '70's. Why? He used aluminum rods with loose clearances, which slung a lot of oil off onto the cylinder walls and cam's underside. This additional bleed-off caused the top end to starve, eventually (in a few seconds) risking a void in supply to the high demand aluminum rods and main caps. By his blocking off the oil through plugs at the back of the block which fed the lifter galley, he forced the surplus to be available for the crank. This was needed because his ability to pump oil through the lower end exceeded the drain back capacity of the top end, and since he was using aluminum rockers with needle bearings, the surplus was also unneeded. Somehow, probably because these "oil restrictor" kits got duplicated and hyped as a "must have," they crossed over into the brainwashed thinking of engine builders of all types of racing. If your engine is leaving too much oil in the top end, revisit your drain back design. If you're using ball fulcrum rocker arms, which needed every edge they can get to pull heat out of that friction point, don't even consider this trick. Don't even consider TAPPETS with OIL HOLES for the cam lobe either. As mentioned above, there is a full 360° spray coming off the side clearance of the crankshaft's rods. It is totally flooding the bottom side of the camshaft, the last thing you need to do, is allow precious pressure and volume within the tappet, that feeds the pushrod and rocker arm, to bleed off on some waste of money gimmick like "cooling" lifters that add oil to the cam lobe. These tappets are a waste of your money.  ^

OPENING THE VALVE: Just as there are two sides of the rocker arm, there are two sides of the camshaft. Most valve train development from three to four decades ago, a pivotal era, had SYMMETRICAL cam ramps. That is, the rates of acceleration to the opening and closing sides of the cam were identical mirrors of either side of the cam lobe's center-line. In the early 70's, ASYMMETRICAL designs significantly increased acceleration of the OPENING ramps, while still setting the lifter down gently on the closing ramps, thus reducing valve bounce and impact shocks. (See: CAM BASICS)

OVER-ARCING and UNDER-ARCING: Is a condition of excessive radial motion across the contact surface of a linear component driving, or being driven by a radial device - as measured by the perpendicular relationship of the radial device's axis to the linear component. In this case, we are talking about the travel of the rocker arm (either side) in relation to the valve or pushrod. Over and Under are opposite consequences, but the net result is the same. In each case the specific angle of rotation from the rocker arm is reduced, compared to MID-LIFT geometry. In addition to this, the acceleration dynamics of the rocker arm upon the valve are influenced, with opposite consequences between the two. "Over-arcing," where the rocker arm is sitting too low in relation to the valve stem tip (the most common condition), will be slower to move the valve off the seat and begin to accelerate as it approaches full lift.  "Under-arcing," where the rocker arm is sitting too high in relation to the valve stem tip, accelerates quicker upon initial opening, then slows down as it reaches full lift, rolling "inward" and UNDER itself. Both of these conditions contribute excessive side loads on the valve guide and valve stem, with subsequent additional heat, additional drag coefficients, increased harmonics in more exaggerated examples and reduced horsepower in all examples. These "symptoms" of bad geometry have often caused engine builders to increase valve guide clearances, change valve specifications and suppliers, as well as other alterations in pursuit of fixing nothing more than "symptoms," instead of the cause (bad geometry). Additionally, dedicated cam testing will be tainted by the lost information from this wasted motion, forcing engine developers to use more aggressive cam profiles to make up for the wasted rocker arm motion; adding more work to the valve train with decreased efficiency and lost power. OVER-ARCING and UNDER-ARCING will impart LESS radial motion to the valve (less valve lift), in addition to accurately mapping the cam's acceleration profile, compared to equally dividing these arcs with MID-LIFT geometry. (See: MID-LIFT ARCING)  ^

OVER-HEAD CAM (OHC): Refers to an engine design whereby the camshaft lays above the cylinder head's valve array, and either opens the valves through direct contact upon them with the cam lobes, or through a leveraged action created by first pushing upon a single or dual end rocker arm, which then opens the valve with direct contact from one end of the rocker arm. In dual end rocker arms with OHC designs, the rocker arm may be above the camshaft, while on single ended rocker arms the cam would mount above the rocker arm. Either single or multiple camshafts may reside upon each cylinder head, controlling one or more bank of valves along the X-Axis (length) of the cylinder head.

OVER-HEAD VALVE (OHV): Refers to the engine design that has the valves over the cylinder block, even though it refers to a "head." This is an ancient and misleading acronym. Acronyms for other valve-train designs, like OHC (over-head cam), or DOHC (Dual-Over-Head Cam) are more accurate in their description. The main point that OHV means, is that the engine design has PUSHRODS and ROCKER ARMS. If you're wondering why Over the "head" is distinguished, it is because the prior design (which predates this author) was the FLATHEAD engines, that had their valves in the block. When the small block Chevy came out in 1955, the term OHV was born; and the Flathead was on borrowed time!  ^

OVERLAP: See: CAM OVERLAP.

PA SERIES: Refers to the STUD MOUNT series of rocker arms that MEI designed for Precision Valve Systems, Inc., back in 1996. "PA" being designated for Precision Aluminum, these rocker arms were made available for Small Block Chevrolet and Ford OHV engines, and Big Block Chevrolet, which also fit (sort of) the Big Block Ford. Taking much of their design innovation from the MEI PRO -STUD rockers, like MID-LIFT Geometry (of course), and a "Precision" (Patent Pending) rocker body silhouette that had its top surface designed to match the mounting stud angle for a "precision" reference of measurement for the engine builder to use in setting "precise" INSTALLED GEOMETRY by using the exact, precise length push-rod. A task made especially simple when used with MEI's Patent Pending G-TooL™. Made from 7075-T6 Aircraft Strength extruded aluminum, the PA Series also incorporates the Rifle Drilled, dual flat 8620 trunnion. As can be seen, the word "Precision" in "PA" does NOT stand for the Aluminum, but the precision engineering of the rocker arm design and means to precisely install it on the engine. PA Series INSTRUCTIONS

PF SERIES: Refers to the MEI designed 4340 chromoly steel, "roller tip" Ball Fulcrum rocker arm, with numerous innovations. Developed in 1996 for Precision Valve Systems, Inc., and provided the unique stamping procedure developed by Jim Miller to allow making a stamped steel chromoly rocker arm, previously unheard of because of the high toll of tooling that chromoly imparted, making them very costly. This led to  the introduction of the OE Series rocker arm, which was aimed at stock rebuild market, but soon became THE "stock type" rocker arm to have, since none other would sustain use in high velocity valve trains (see: OE Series). The first feature that made this rocker a must have for a high strength roller tip design, was the 4340 Chromoly steel which no other stamped roller tip rocker had. The other concept of the novel design to such a rocker was the MID-LIFT roller tip rocker principles. The rocker arm also incorporates the precisely ground flat top surface, which facilitates a reference plane for measuring with the rocker stud to get exact Installed Geometry, especially when used with MEI's Patent Pending G-TooL™.  The result, like any MEI MID-LIFT geometry rocker arm, is that the area-under-the-curve was at its greatest, giving the most valve lift, velocity and duration for each degree of crank motion, throughout the entire valve lift cycle; while showing the same, minimal back-and-forth sweep atop the valve tip. The in-and-out pushrod motion was also nearly non-existent. PF Series INSTRUCTIONS

PS SERIES: Refers to the PRO-STAND™ "shaft" style rocker system developed by MEI in 2003 for the most popular brand and model aftermarket and OEM high performance OHV cylinder heads. It was developed to combine our novel manufacturing technique previously limited to NASCAR style rockers from Billet aluminum, with unified rocker offsets as needed per application so that a new standard of Professional, high strength, MID-LIFT geometry billet rocker arm and stand could be sold for a near "sportsman" level price. This approach of manufacturing places the molecular grain of the alloy "in--line" with the rocker body loads, rather than cross-grain as with the "extruded silhouette" designs used by all other "name brand" rocker manufacturers. We also have infinite flexibility in dimensional changes and material use so the very lightest while being the very strongest design can be assured for the amount of material used. Custom designs are also possible from concept to finished product in 24 hours! The instructions for the PS Series are unlike any other illustrations shown here, in that they were purposely designed and published so the ZOOM feature on your Adobe browser will allow you to zoom in for a high resolution, to see the detail of minimal motion results on top of the valve, in matching results of your installation.  PS Series INSTRUCTIONS  For current Product Line see:  PS Series APPLICATIONS

PEAK FLOW (Head Lingo): Refers to air flow through a cylinder head's intake or exhaust port. Specifically it is the point of valve lift where, as measured on a flow bench, the proportion of increase in air flow for the proportional increase of valve lift, as measured in "area," converge and go negative. RULE OF THUMB: When the increased value of air flow per 0.050" increase in valve lift reduces to less than 10% increase in CFM, the port is near death; and increasing valve lift beyond THIS POINT (NOT the "peak" where it loses air flow), is unnecessary forces on the valve train, and thus lost horsepower.

PEDESTAL: This refers to the rocker arm mounting, usually of OEM design which incorporate a fulcrum that has a base to it which provides a solid mount directly to the head, whereas it uses a bolt instead of the older stud and nut design. The bolt is mounted directly through the pedestal/fulcrum and tightened until bottoming out. These are usually "fixed," non-adjustable valve trains where all valve stem length dimensions, cylinder head deck heights, etc., must be set to factory specs. The hydraulic lifter is the only variable for minor assembly length variations of the valve train; and this variable is typically only .050".

PHASING: Is a term used in engine building which pertains to the synchronizing of the camshaft with the crankshaft; more specifically, the cam's INTAKE LOBE with the crankshaft's piston position within the cylinder, measured in crank degrees of rotation relative to the exact TDC (Top Dead Center) point of piston position. This measurement of the cam is often done in one of two ways, depending on the engine builder's preference. One, is by using the intake cam lobe's centerline for reference to the crank position; or two, the actual opening points on the cam ramps are checked with the cam card, dominantly chosen as the .050" lift position of the lifter. A dial indicator mounted and aligned precisely parallel with the tappet's motion is essential for measuring the cam motion, while a large degree wheel that has been accurately synchronized with the piston to indicate exact piston TDC is also needed. Advancing and retarding the cam from the cam manufacturer's published specifications is done after establishing accurate phasing first. A general analysis of a cam's acceleration rates can be done by using graph paper to plot the tappet rise and fall in 5° increments, but for precise analysis of the cam lobe, more sophisticated instruments with computer software are available to check the cam on a bench fixture or in the engine. But "phasing" is a term limited to simply timing the cam with the crank.

PISTON to VALVE CLEARANCE (PV): See: VALVE CLEARANCE.  ^

PISTON VELOCITY: This is a relatively self descriptive term, but its real influence with engine performance is often not fully understood. Piston velocity, or "speed," is the result of two things: (1) Crankshaft stroke, and (2) RPMs (obviously). Piston speed is THE SOURCE for response characteristics of cylinder head air flow through the ports, in addition to the cubic inch volume each cylinder has with every stroke. These are two crucial elements to place in proper context: (a) Airflow VOLUME, and (b) airflow VELOCITY. Cylinder volume in conjunction with piston speed determines all of the tuning characteristics and specification needs within a given cylinder head's design, as well as the fundamental camshaft specs to meet these needs. These two elements can offset each other when choosing cylinder head packages and/or camshafts. In other words, if you have a large cubic inch small block engine, its increased stroke will achieve high torque at lower RPMs compared to a conventional size small block engine, but it can use larger port heads and greater camshaft specs that would more typically be chosen for a high winding engine of smaller cubic inches.

NOTE: Torque is often the term that is used to explain what a large stroke engine sees. Although it is true that the operating pressures created with each power stroke is applying leverage onto a longer "arm" for an engine having a greater stroke, and thus "torque" forces upon the crank are increased, the simple truth with what is happening in the cylinder heads with airflow is that TORQUE is a symptom from the increased PISTON VELOCITY that the stroke generated. The engine literally thinks it is going 8,000 RPM when the tach reads 6,800 RPM, because airflow and volume demand are greater at lower RPMs. There is a second dynamic to the piston's influence on the cylinder heads, dictated by a ratio of connecting rod length divided by crankshaft stroke. Known as the ROD RATIO, this determines how quickly the STARTING and STOPPING of PISTON SPEED occurs. But the ultimate piston speed is still determined by our two main factors (crank stroke and RPMs). The formula for determining piston velocity in FPM (feet per minute) is this: STROKE x 3.1416 x RPM ÷ 12 = FPM. If you use this formula when comparing different engine strokes, you can see how the RPMs are affected. TO COMPARE, pick your first stroke combination and use the foregoing formula. Then save the FPM value in your calculator and begin working backwards from this sum. Only now substitute the division symbol for a multiplication symbol, and divide where there is a multiplication symbol. SKIP over the "RPM" in this second step working backwards while changing the STROKE value to your second choice (i.e., 4.50" first calculation compared to 4.00" for second comparison), and your final answer will be the different RPM "effect" created in the ports by the different stroke. The answer to a comparison of a 4.50" stroke running at 6,000 RPM, compared to a 4.00" stroke is 6,750 RPM. In other words two identical engines of the SAME SIZE cubic inches having their only difference being stroke, will have a 750 RPM increase being needed for the 4.00" stroke to have the same port velocity as the 4.50" stroke. Or more simply stated: the 4.50" stroke engine only needs to turn 6,000 RPM to make the same power that the 4.00" stroke engine can make at 6,750 RPM. When you consider that the stroked engine will also have more cubic inches (unless the cylinder bores were much smaller to keep the comparison the same size), the extra VOLUME dictates that PORT passages would also need to be increased, as well as VALVE LIFT. This is why it is easy to exceed the cylinder head capabilities when going into big stroke engines; not to mention all the increased work needed by larger induction and exhaust. When these limits are reached with such engines, the RPM's stop dead in their tracks, although you've created a "torque monster" at lower RPM's; and to think, it all starts with "piston velocity."  ^

PIVOT HEIGHT: This is a term we refer to define the height of the roller pin axis above the trunnion/shaft centerline, when the valve is closed, which equals exactly half of the NET Valve Lift.

PIVOT POINT: The axis of the trunnion or shaft in defining the rocker body design. But also refers to the roller tip, when describing the height of the trunnion/shaft when pivoting the entire rocker body up or down from the roller tip, to establish INSTALLED GEOMETRY, as noted above in pivot height.

PLENUM CHAMBER: (Induction Systems) A component of the induction system which is designed with or added to the runners of an intake manifold. It resides on the entrance side of the intake ports, adjoining the runners into this common open chamber of volume, which is usually beneath the fuel delivery, whether it is a carburetor or fuel injection system, although some designs of fuel injection systems with injectors having direct cylinder head mounting have been used with plenum chambers atop their runners. Usually however, individual runner designs with a single throttle body (butterfly) or independent throttle valves per runner are chosen. Its main purpose is two-fold: (1) To provide a "buffer" between the REVERSION waves coming back from the reflected flow hitting the closing intake valves, and avoid their disturbance of the fuel metering of the carburetor or injector's metering valve (on non-direct port injection designs). (2) Provide a larger "bank account" of air volume directly above the operating port runner that demands it; drawing from the collective supply of air flow through multiple butterflies that usually feed the plenum chamber, which otherwise do not offer adequate volume individually for the instant, millisecond demands that each cylinder would make at high RPMs.

NOTE: When these reversion waves have not been arrested through a tuned length or a plenum chamber, they reach a frequency at high RPM's which actually escapes through the induction system's entrance; which in carburetors and some fuel injection systems creates a condition known as "stand off." The phenomena can be seen as a fog or standing mist above the carburetor inlets and destroys power.

POINT-OF-INSTANT-CENTER: Refers to the tangent point where two intersecting lines meet, usually mentioned with angular devices which pivot and form linkage where this pivotal tangent point occurs. It is a phantom axis point. Pushrods that lean into the angle established by the valve centerline create a point of instant center several inches above the cylinder head which determines load force direction (induced by resistance to that force) upon the mounting locations of the rocker arm. A more appropriate use of the term applies to CHASSIS SUSPENSIONS, where 3-Link, 4-Link and 5-Link control arms, ladder bars and other similar devices mount in a way that run parallel to each other on one plane, with a positive attack angle on another plane that eventually creates a tangent point in space beyond the lengths of the component linkage itself, and having the effect as if they are this total length. A four-link suspension system on a drag car is adjusted so that if a line was projected out from both the upper and lower links, where they meet will have the same force of lift as if a more traditional ladder bar was used and its forward mounting point was at the same tangent point.  ^

POLY LOCK: See: ADJUSTING NUT. (Nick name or trade name traditionally used in slang for ADJUSTER.)

PORT VOLUME: Cylinder head lingo. This is usually referred to in cc's (cubic centimeters) of space. This is something that is nearly ridiculous to real cylinder head evaluation for port flow quality. It is a value, only relative to tuning or cylinder head and manifold design aimed at specific cubic inch engine sizes, where plenum volume in the intake manifold, rpm's being tuned for and engine sizes are all taken into consideration; and usually requiring computer analysis. Yet, this is often used to SELL cylinder heads, or EVALUATE cylinder heads, and this is NOT a good barometer for either. The only exception to this, is when the cc port volume of one given cylinder head combination is used against another by an engine builder who's quantified the "package" under a specific engine combination. This method of measuring a head is only good for calculated ram tuning 'theory." Often with computers. It has no sense in reality to really judging a good head from a bad head. For example, with a high speed hand grinder I can concentrate my material removal in one area of a port, thus increasing its "volume," while doing nothing to the valve bowl area, doing nothing to the intake port entrance, or do anything to the greatest bottleneck of the port (wherever that may be). But I have increased cc's, and the port is "bigger" by that measure. However, the FLOW BENCH (and Dyno) is going to show "turbulence" from the change in velocity as the air moves from a smaller cross sectional area of the port and into my new "ground out" area, where it will slow and tumble. Air flow "quality" is going to go down. Velocity is going to drop (from the turbulence), BUT... I am going to have a LARGER "cc" port. Another analogy to show how illogical this is, would be to compare a Ford head with a LONGER port, to a Chevrolet head with a SHORTER port, and of course the LONGER Ford head will hold more cc's of volume, so theoretically it is supposed to be "bigger?" I don't think so. But that's the goofiness that some companies and distributors try to sell heads by, or some gurus try to promote "tests" by. Forget "cc's!"

PORT VELOCITY: (Cylinder Heads) Refers to the speed of air flow traveling through any given cylinder head's intake and exhaust ports; usually with more emphasis on the intake ports. This is due to the entirely different relationship between relatively cool intake port airflow  versus extremely hot exhaust port airflow, which is really beyond using "air flow" as a term to describe violent high pressure gases escaping from the cylinder. With regard to the intake port, Port Velocity is the result of the Piston Velocity's influence on the port's cross sectional area. Even though its final value is determined by several other factors, such as cam timing, port length, reversion pressures and vacuums, on a naturally aspirated engine it is still ultimately limited to these two, fundamental components: CROSS SECTIONAL AREA and PISTON VELOCITY.

PORT WINDOW: Used in Cylinder Head lingo, refers to the view one has down the intake or exhaust ports of a given cylinder head, where maximum view of the valve head is seen. See: ROCKER MOUNTING & VALVE TIP HEIGHT for additional illustrations that depict port silhouettes within the head, relative to valves.

PRE-IGNITION: (Cylinder Heads) A term used to describe the untimely ignition of fuel in the cylinder before the timed ignition by the spark plug; usually a result of white hot residual particles left over from the previous cycle, or an excessively hot valve, resulting from too lean or inadequate fuel mixture, and/or improper timing which increased operating temperatures. See: FLAME PROPAGATION.

PRECISION STAND™: A term was dropped by MEI when summer 2005 development on the PRO-STAND systems was consolidated into a new level of high end rocker systems at affordable prices.

PRIVATE LABEL: A term used within the automotive aftermarket trade that refers to the sale of name brand products under a generic name, usually to a peer company or distributor who doesn't want the trade name used. Many companies buy products from name brand manufacturers, then repackage them, or have them packaged with their brand name on either the product and/or packaging.

PRO-SHAFT®: Is a Registered Trademark ® of MILLER ENGINEERING INC (MEI), Pompano Beach, FL., and the original, premium level STAND mounted rocker group for MEI, which originated from the BOSS 429, on through NASCAR designs of the mid to late 1990s.

PRO-STAND™: A Trademark Pending name for MEI, it is a newly released product group pertaining to STAND mounted rocker arms which has overtaken two design groups for two markets. PRO-STAND systems represents the ultimate, high end rocker arm for price conscious engine builders, that is also designed for distribution through select, top WD suppliers. Without a doubt, the ultimate "bang for the buck." PS Series INSTRUCTIONS   ^

PRO-STUD™: Is a pending trademark and a product group rocker arm of MILLER ENGINEERING INC, designed for maximum performance STUD mounted valve train requiring minimum weight, ultimate strength in an aluminum body rocker arm.

PUSH-ROD: The link between the CAM FOLLOWER and the ROCKER ARM; usually made of a single piece of steel tubing between 5/16" diameter on through to 7/16", they have many variations of length and end designs made to accommodate either a male or female radius end connection on both ends to provide freedom of misaligned movement while operating in a linear path. They have been made in different and exotic alloys, as well as shapes and end design combinations; but the net result and purpose is still the same: push the rocker arm up with the least amount of flexing.

PUSHROD CUP: The heat treated, pressed-in steel component of the aluminum rocker arm which mates with the ball-tip of the pushrod.  ^

PUSHROD GEOMETRY: This is a term we've used to be specific in explaining that "rocker geometry" is based on "push-rod" angles with the valve. They are one in the same. See: GEOMETRY, DESIGN.

PUSH-ROD HARMONICS: This is the vibratory motion that occurs at different frequencies and RPM'S from the side-to-side motion, the rebound of compressed forces, influences from heat, friction and worse: high performance operational linear loads on a non-linear axis. In simpler terms: the pushrod misaligning with the tappet's centerline, and excessive in-and-out motion following a pushrod cup (or adjusting screw) around the rocker's axis more than it needs to. Like a worn tire on a car that shakes at certain speeds, but smoothes out at others, "harmonics" also occur in waves before reaching their terminal point. Inaccurate and over-arcing rocker arm geometry magnifies both their symptoms and effects, which we believe also dominos into early valve spring harmonics that then tend to feed each other. See: HARMONICS.

PUSH-ROD LENGTH: This dimension is critical. Accurate pushrod length is also very under estimated and misunderstood as to just how critical it really is. The fact that pushrods are sold by nearly everyone in .050" increments is testimony to how casually the valve train industry treats this; when a fraction of this amount can influence area-under-the-curve and lost cam duration by several degrees. This is the essence of INSTALLED GEOMETRY; especially on the stud mounted rocker arms. Pushrods are the LAST components that should be bought for an engine, since their length cannot easily be established until the full valve train has been mocked up with the cam, lifters and finished cylinder heads in the final machined engine block. On stud mounted engines the pushrod length provides the sitting height of the trunnion below the roller tip's axis when the valve is closed. This "sitting height" is measured in angles to the valve stem centerline by a line that runs between the trunnion axis and the roller pin's axis. Shortening the pushrod drops the tail of the rocker arm, increasing this angle and lengthening the pushrod raises the trunnion on the stud and decreases this angle. This angle is always going to be HALF of the total angular motion that the rocker arm will rotate for its given valve lift. (See: FORMULA) Very few engine builders will need to know, or bother with converting dimensions to know these angles, especially to set their installed geometry accurately, since we instruct and provide easier alternatives. This is simply done by measuring the height of the trunnion's centerline below the roller pin's centerline, and adjusting the rocker's height to be half of the net valve lift. (See: INSTALLED GEOMETRY for illustration.) Because the dimension which influences this angle comes from the trunnion, but our adjustment comes from the pushrod, a required changed in trunnion height of .050" will require changing the pushrod by approximately whatever the rocker ratio is; which would be .075" for a 1.5:1 rocker arm.  ^

REASONS: For those who may have missed the explanation given in The History Of MID-LIFT, as to why pushrods NEED changing when factory spec rockers are changed to roller rocker arms, even if the valve lift stays the same, here is the summary: Roller tip rocker arms have "roller diameters" that vary from .480" to over .600". Factory rockers are "shoe" tip designs. The measuring point for what influences rocker geometry on the valve tip, is measured at the centerline of the roller, where in contrast, it is measured at the contact point on the bottom of the shoe tip design. Simply substituting shoe tip for roller tip ELEVATES this measuring point by HALF of the roller's diameter. So all else being equal, and assuming the factory rockers were anywhere close to correct for the valve lift, you would have to raise the rocker arm UP by HALF of the .480" diameter roller, or .240". But remember, this is AT THE TRUNNION. Which means you would need a PUSHROD LENGTH that was 1.5x LONGER than that, or .360" Longer than stock! If you've increased the valve lift by changing cams, or increasing rocker ratio (or both), then the MID-LIFT POINT of valve travel will be LOWER than stock valve lift (further from the valve tip, downward), so this increased trunnion height (rocker height) will be lower too, thus offsetting this .240" dimension and .360" longer pushrod. If valve lift was increased .100", then the MID-LIFT point would be LOWERED by .050", so the .240" trunnion height over stock would now be only .190" and the increased length of the pushrod would only be .285" (don't forget, it is changed 1.5x the trunnion height dimensions, so the value would adjust .075" less).

NOTE: Historically, because installed geometry has been unknown of before the principles of MID-LIFT were developed, it's importance has been ignored by the valve train industry, so it is not uncommon to find that most rocker arm mounting studs are too short to provide an adequate number of threads above the rocker arm, once pushrod length has been increased for MID-LIFT geometry. The desired minimum threads above the trunnion should be 1.5X the diameter of stud. (i.e., 3/8" stud= 3/8" + HALF, or .562" height of threads above trunnion; 7/16" stud = 1.5x .437", or .655"). The very minimum should be no less than equal the stud's diameter. On STAND (shaft) MOUNTED rocker systems the pushrod length is just as critical, except rocker height adjustment is established by the stand itself; often machining its bottom to reduce rocker height, or "shimming" it UP to raise it. The adjusting screw on stand mounted rockers is kept to a specific length within the rocker body, usually determined by how many threads "out" from the cup being fully bottomed out within the rocker. Jim Miller established this standard as "two threads out" back in 1978, and has pretty much been adopted by the more well known "shaft" rocker system companies since. (Current MEI design STAND mounted rocker systems use very specific overall length adjusting screws, flat washers and NUTS to control this fixed dimension even more precisely, which anticipates several different needs in setting geometry and tuning with lash adjustment, setting another precedent.) Once the stand has been set to its necessary height, and the adjusting screws are set to two threads out, an adjustable pushrod is then used to zero out between the lifter and the rocker, PLUS valve lash is set, before a final overall measurement is taken to then make the exact length required. There is NO ".050" length averages used in buying pushrods here. At this level of investment and sought after performance, the pushrod length is ordered within .005" of exactly what is needed for MID-LIFT precision.  ^

QUICK LIFT: With regard to rocker arm geometry, it is a term recently applied by a well known cam company to try explaining some logic to over-arcing rocker arms, which is not explainable (but that's just our opinion). See: VARIABLE RATIO.

RADIAL ARC: As we define and use with rocker arms, technically, is defined as "developing symmetrically about a central point," which is about as accurate a description of the MID-LIFT principle as it can get. Because it is always (and only) measured perpendicular to the axis of rotation, its lateral motion (in-and-out sweep) is the ultimate minimum.

RADIAL MOTION: Meaning "radius" (or circular motion), is the general measurement of any circular motion from a specific radius, but NOT relative to any specific lateral effect (sweep) or angle of incidence (attack angle). It's a generic term used in a general way to convey the difference from, and relative to LINEAR MOTION.

RAM TUNING: Induction system lingo. This refers to the design and tuning of an engine's induction system, and exhaust system, to achieve a flow depression across the piston that exceeds the standard vacuum it creates through natural operation. It is achieved with a concerted effort of exhaust manifold lengths, collector design and often cross-over tubes for the "initiating" aspect of creating this cylinder depression. But the cam and intake manifold are developed to compliment and complete the concept so a precise volume of air is waiting to fill the preceding and magnified flow depression in the cylinder that was created by timing the exhaust's supersonic reversion waves with: (1) the exact moment of peak piston depression and (2) cam initiation for the subsequent cycle. This is designed for a specific RPM for all this to occur. In simple terms: the vacuum waves created by the exhaust system are synchronized with the FLOW DEPRESSION created by both the piston and the reversion waves of the intake tract, so that inertia of the following volume of air (measured by port volume) can over-fill the cylinder. Sort of a naturally aspirated super-charging. See: Volumetric Efficiency.

RATIO: Referring to rocker arms, is the difference in lengths between the rocker's axis of rotation and its two respective opposite ends, divided into each other. This is broken into TWO specific divisions of terms: MECHANICAL RATIO and NET RATIO. The mechanical ratio is the "paper" and actual measured ratio taken between these three points; while the NET ratio is the actual, multiplied "running" ratio imposed on the valve by the rocker arm AFTER FLEX, MISALIGNMENT, GEOMETRY MOTION and other real world factors have been included. All MEI engineered rocker arms are designed for NET RATIO at ZERO LASH, whereby on each engine we have already calculated a compensation factor into the MECHANICAL ratio to allow for these variables. Each engine is unique to these required compensation factors, and any deviations from MID-LIFT geometry will also change the NET ratio.

NOTE: Throughout the history of roller tip rocker arms, no geometry standard was ever established, or even attempted by rocker arm manufacturers or cam companies. This is a fact. Rocker ratios varied greatly between "advertised" and what the engine builder actually ended up with. It is important to remember: RATIO is only a SNAPSHOT taken at FULL LIFT of the valve. A rocker arm can easily show an accurate ratio when measured at full valve lift, but lose valve LIFT, DURATION and ACCELERATION throughout the rest of the lift curve, because of wasted motion. Rockers were traditionally designed from a "closed valve" perspective, as if it would never open. Old rocker drawings of top companies illustrate dimensioning in this position, where it is easy to see why the rocker's length was designed too long. The pushrod cup was too high, so in operation the rocker would follow an excessive arc up and in toward the stud, losing NET ratio at the valve. Flex was incorrectly blamed for this. In the late 1970's, some manufacturers started adding a correction factor, by moving the adjusting screw (or pushrod cup) closer to the trunnion to increase the mechanical (paper) ratio, until they eventually achieved a NET ratio that was close to advertised. But the over-arcing pattern of the rocker's DESIGN GEOMETRY still existed, and this resulted in slow response at the valve throughout the other points of the valve lift cycle before reaching full lift. The consequence was LESS AREA-UNDER-THE-CURVE, which required additional cam rotation to lift the valve the same amount as a MID-LIFT rocker moving the valve to the same point of lift, wherever that measuring point would be.  ^

RATIO DIFFERENCES: From time to time I am asked why we don't offer exactly the same ratios for some applications, like the 1.73 Ford ratios, and instead we only offer the 1.70. The short answer is the market doesn't demand it. But, even though it overlaps some of the points made above, my long answer is below:

1. As noted above and elsewhere on this web site, we've often clarified that "ratio" is a "snap-shot" at full lift only, which does not account for the accurate ratio of any rocker arm throughout its arcing cycle; which is a different dynamic that affects cam performance (at the valve) drastically, and is known as "area-under-the-curve." Also, because "geometry" of SHOE rockers is entirely different in the way the rocker arm responds to SPEED and DURATION at given points in between CLOSED valve and OPEN valve - when RATIO is determined against roller tip rocker arms. So we don't need to "nit-pick" with our rockers in trying to duplicate arbitrary ratios that the OEM factories came up with.

2. You will also see that there is NO STANDARD between any company's design and another, because before MID-LIFT was developed, all companies chose random CLOSED valve relationships, VALVE lengths and all designed their rocker arms from the CLOSED valve position, again, with NO REGARD for the various ratios that the rocker arm goes through BETWEEN CLOSED valve and OPEN valve.

3. If you check other companies, you will see their NET (real) ratio, including Ford, Chevrolet and GM, are not exactly accurate to what is "published" or specified as the "design" ratio. None of these companies, ever specified a standard of any accuracy, nor did they educate anyone as to what their ideal settings of rocker heights for GEOMETRY was, until after MID-LIFT was published; then studied; then adapted by them. Now it's the only geometry that Chevrolet or Chrysler recommend. What difference does this make? Well, if you change the PIVOT height of any rocker arm when checking ratios, you'll notice that the NET RATIO CHANGED also. There is no exact value for what ratio will be, WITHOUT a known geometry value.

4. The ratios were often selected NOT for any engineering reason, but to "match parts." There were no engineers who found a secret to anything in the valve train that established a  "1.73" ratio was better than "1.70 or 1.75." Usually oddball ratios are derived by a stack of linear, compounded measurements of all the components from the cam centerline up through the tappet, then the cylinder head mounting points for the rocker studs, to the valve centerlines... and... while trying to make as many of these parts fit EXISTING and FUTURE engines. Hence, the same rocker arm was used for BOTH the 351 Cleveland and the 460 Ford, while the ROCKER and entire VALVE TRAIN GEOMETRY between each of these engines is as different as night and day! See: FORD'S PRECEDENT

5. Lastly, "we" choose ratios of High Performance rocker arms, aimed at high performance engines, to be selected and used by high performance engine builders, who are usually building for high performance applications, who have to assemble a mix of manufacturer's parts. So there is no incentive for us to copy any known value established by any NON-high performance application, especially when the company who designed the original engine made such compromising choices just to save a few dollars.

As stated elsewhere here, if you were to measure the amount of CRANK rotation that it takes to open the valve with any MID-LIFT rocker arm, properly installed, you would find that MORE VALVE LIFT occurs THROUGHOUT the valve lift cycle from LESS CRANK rotation (high efficiency) than any other rocker arm design, especially factory SHOE tip designs, which have an entirely different TRACKING VELOCITY than a roller tip rocker, regardless of the manufacturer.  ^

RATIO, INCREASED ROCKER RATIO benefits is a question we are often asked. "Why do different engines use different ratios?" "What is the benefit of more ratio?" "What is the consequence of increased ratio?" Here is a few points to understand and consider. In the beginning, as with many engineering ideas pouring in from all over the world, there was no standard, simple answer. There was no need for one. Rocker ratio's are an easy way to increase the needed movement at the valve with less movement at the cam. These were spawned by physical space limitations in engine blocks and obvious benefits to efficiency in multiplying this motion, as engineers realized greater valve lifts over the decades of engine evolution. It was also easier to experiment with valve lift and timing through simpler changes in rocker arms atop the engine, rather than redesigning the block, the cam, and so on as engineers tried increase the engine's power. (This was, by the way, spawned heavily by aircraft engine development during and after WW1.) The 1.50:1 ratio became a good balance many engineers settled on, making slide rule calculations (literally) an easy denominator to deal with. Ford, Chrysler (and others) found a need to adjust this a little more, increasing up to factory OEM ratios in the mid-1.70's. But the reality is, most ratios were chosen for convenience in keeping with traditional values that worked with existing cams, where no real revolution to change was seen.

Keep in mind, that the effects of valve lift acceleration from a roller tip rocker arm is entirely different from that of a friction pad "shoe tip" rocker arm. Because of this, much of the aftermarket industry's influence with roller tip rockers on modified Detroit passenger car engines was misinterpreted, even though they would be advertised, sold and used as "stock ratio." (i.e., 1.50:1 for small block Chevy.) Why does this matter?

As an engine builder choosing and buying an assortment of components, probably from a variety of manufacturers, you need to understand a little bit about the history of these parts to make sense of the conflicting sales hype you may hear between the different companies which compete with each other. Most cam companies sell rocker arms. Some rocker arm companies sell cams. So of course they each want to sell you their companion parts, but as a savvy engine builder, you know that some cam companies don't make the best rocker arms, and vice se versa. Go to a mail order warehouse, and you'll need your lunch to listen to all the hype (sometimes). Most important to the above points, is understanding what stage you are at when changing rocker arm designs (roller tip vs. shoe tip), and existing NET ratio to your possible increased ratio. LASTLY and MOST IMPORTANTLY to consider, is the GEOMETRY of the rocker arm you've been using, compared to where you are about to go. This applies to both DESIGN and INSTALLED rocker arm geometry. (You're not surprised I would mention this, are you?) You can choose a 1.50:1 MID-LIFT rocker, and easily see as much valve lift per degrees of crank rotation throughout the lift cycle, as possibly using a competing brand of rocker that is advertised as 1.60:1.

But here's the short version to the main questions about INCREASING ROCKER RATIO. All else being the same (roller tip for roller tip, shoe tip for shoe tip, etc.), increasing rocker ratio will increase ALL THREE dynamics: VALVE LIFT, VALVE DURATION (note that I did NOT say "cam duration"), and RATE of ACCELERATION (velocity). Bear in mind it will do it on BOTH sides of the cam cycle, OPENING and CLOSING. In "some" cases, this quicker landing of the valve upon the seat from a quicker rocker ratio can increase wear. Of course the most obvious warning is PISTON to VALVE CLEARANCE; although in most cases, "moderate" increases in ratio (i.e., 1.50 to 1.60) won't often create a problem, if there was a surplus to begin with. But YOU MUST CHECK to be sure.  ^

There's always been a trend to "split ratios" on many engines, where a higher 1.60:1 intake rocker ratio may be chosen, then accompanied with a 1.50:1 exhaust rocker ratio. It is not necessary or even beneficial to do this when you are building from the ground up, and not mixing parts you need with parts you have. I've commented on this elsewhere, but setting the rocker arm's INSTALLED geometry with a different pivot height to the valve tip (in other words, how low beneath the roller tip's axis the rocker shaft's axis sits) should be done with changing the height of the valve tips themselves. Since the exhaust valve on most applications will have less valve lift, its rocker arm's pivot point (shaft) would sit higher in relation  to the valve tip. Or more accurately stated, the exhaust valve tip would sit lower compared to the intake valve tip. This pertains to STAND (shaft) mounted rocker systems, where a common shaft may connect the intake and exhaust rockers for a given cylinder. On individually mounted stands, such as the big block Chevy might have, this doesn't matter. You can match all valve tips to the same height, and designate your stand height for each, as needed to achieve MID-LIFT geometry. Rocker ratios should remain the same, for several simplicity reasons to design and use, and the differing specs needed between the intake and exhaust valve should be ground specifically into the camshaft. Don't use the rocker arm for a "tuning" tool.

THE REAL BENEFIT to increasing rocker ratio, is that it allows for a larger base circle camshaft, which on a full competition engine, is often a good benefit. The rotational torque, flex and "whip" imposed on a camshaft operating with high spring pressures at high RPM's is horrendous. Increasing the cam's base circle aids the cam engineer in adding precious material where it resists this, and also in the cam design itself, there is more room for creating a "fast" ramp (more velocity). There is a balance to this though. The rocker arm must be designed to withstand the additional beam loading without adding more spring board effect to the rebounding that this extra pressure induces. So too must the pushrods, tappets and all mounting hardware be adequately beefed up to accommodate the extra pressure that is proportionately increased with the percentage of increased ratio.

It should be noted that only in flat tappet camshafts does this increased ratio give the most relief to the cam's base circle, but because a flat tappet cam is a "friction" component, rubbing directly atop the cam lobe, it also suffers the greatest limitations to increased valve spring pressures. Roller tappet profiles have no real mechanical limits to how fast the tappet (and valve) can be accelerated, as that imposed by a flat tappet profile. But of course, there ARE limits to inertia, mass (weight) and velocity which are cut in stone with specific material designations and designs. For your information, we've designed and tested (successfully), as have other rocker arm manufacturers, ratios from standard 1.50:1 to over 2.20:1 for NASCAR designated engines. But the cams were specifically and carefully designed for these. When all was said and done, everyone came back to earth and still relied on designs between 1.70 and 1.90.  Even though some may have found 1 horsepower here or there stepping outside of this, there was always a counter-benefit that offset the gain. Much like life.  ^

REVERSION: (Cylinder head lingo) Is a phenomena of reflected airflow which occurs in both the intake track and exhaust passages of an engine, predominantly in or near the cylinder head, but with the induction system, often within the intake manifold (or injection ports), and up through the carburetor or metering valve. It is the by-product of closing valves on the intake side, and subsonic and supersonic displacement of mass within the exhaust track. With the INDUCTION side, reversion waves are created within the intake ports by the reflected air that slams against the closing intake valve and bounces back up the intake passages. Proper runner length (and diameter) minimizes this by more evenly matching the VOLUME of air demanded by the cylinder; which is accomplished by the amount of volume contained in the runner and adjacent PLENUM CHAMBER if one is used. With the EXHAUST side, reversion is instigated and dynamic in a different way. More simply put, when a high speed exhaust wave (known as the primary wave) penetrates the existing atmosphere within an exhaust track, it creates a vacuum behind it that rushes in to fill it, which travels in "reverse" back up the path just made by the primary wave, which we refer to as the primary reversion wave. Limiting the reversion wave's entrance into the exhaust port of the cylinder head through various means, including internal baffles within an expanded tube of the header flange, have had some success on the Dyno. Port matching of the exhaust header tube with the cylinder head ended decades ago, opting for a deliberate step or "wall" that the reversion wave would see as it reaches the cylinder head's exhaust flange.

NOTE: The precise use of specific length exhaust tubes creates a timing of these waves within the cylinder which provide an instant low pressure atmosphere to aid in the succeeding cycle's exhaust process, but this is only tunable to specific RPMs. When done, optimum Volumetric Efficiency can be achieved beyond 100%. It might be worth noting that we call this the "primary" reversion wave because there is a "secondary," and a third and a few more, depending on how bad the cam timing and port (and tube) dimensions are at a specific RPM of operation. These high VE results only come at specific frequencies when these subsequent waves reach what might be called a balanced frequency of the primary wave. At other points in the RPM curve, they can actually inhibit each other.  ^

RETAINER: (Cylinder Heads) See: VALVE SPRING RETAINER.

ROCKER ARM: Mounted atop the cylinder head in a variety of different ways, but always with one end atop the valve stem tip; the rocker arm is a "radial" instrument, designed to convey and transfer the "linear" movements of the pushrod and the valve, respectively. There are TWO (2) sides to the rocker arm, of which both sides "rock" back and forth about a common axis, usually pivoting on a "shaft," a "trunnion," or a "fulcrum." Yet, arguably, for more than 34 years of aftermarket engineering, it has been designed as though the valve is the only component to move, with well known companies teaching very bad and inaccurate philosophies to unsuspecting engine builders. One such fallacy, is to "use longer pushrods" to "place the roller in the middle of the valve stem" tip. Other trinkets have appeared over the years, supposedly to check this necessary pushrod length and set rocker geometry for you. But as you read a little further, you will notice there are many variables to setting a rocker in its correct location.

ROCKER DESIGNS: Every rocker manufacturer has their own reasons for any particular design of a rocker arm - but the truth of the matter is, most copy an already well used design; or worse, take what they think are good ideas from two or more designs and put them together. Without getting into mud throwing, which is not the intention of this information, let me simply state: there are so many variations between one company's rocker arms and another - for the same engine application, that using the logic of placing the roller on the center of the valve, will - more often than not - place the running geometry in a different state for each brand. I've even seen the SAME BRAND of rocker arm for the SAME ENGINE application have different lengths between the trunnion and the roller as a way to change the ratio.

NOTE: On a more fundamental comparison about rocker arm design, I should point out something that I often have to explain to people when they compare going from a "shoe tip" rocker arm (OEM) to a "roller tip" rocker arm: The trace geometry (path of tangent points) for a shoe tip rocker is entirely different as it goes through its arcing process to open the valve. With a shoe tip rocker arm, the contact point between the shoe and the valve tip is actually stretching and changing its effective length. How much change is determined by its installed geometry. You can actually mitigate (reduce) how much deviation there is, or increase it, simply by playing games with the rocker's pivot point height with the valve tip. But you're also changing how much and how fast the cam information gets to the valve, at various points of the valve lift curve. With a roller tip rocker arm, this arcing motion is constrained to the "axis" of the roller tip, which rolls back and forth, but doesn't really stretch its length as a shoe tip does. The comparison of these two dynamics is easy to see side-by-side, but is complicated to compare in a precision manner without dedicated cam and motion analysis tools and software. Let's just emphasize that the "rate" of change at the valve will be much different between a shoe tip rocker opening the valve, and that of a roller tip rocker on the same engine. This is also why I have always scoffed at cam or rocker companies horsepower claims being attributed to solely a change to their rocker arm. Even if a horsepower increase occurred, it was most likely the result of the valve motion seeing a bigger cam simply because of geometry change, and not from "roller bearings" or even an engineering virtue to the rocker brand over another, although such differences surely exist. Changing pushrod length to adjust rocker geometry when going from a shoe tip rocker to a roller tip rocker, all else being equal, is mainly for this shift in measuring points that a roller rocker has over that of a shoe tip design.

ROCKER GEOMETRY: There are TWO kinds of rocker arm geometry: (1) DESIGN geometry and (2) INSTALLED geometry. You can't do too much about "design geometry" because that is determined by the manufacturer of the rocker arm. But the "installed geometry" is the result of how you installed the rocker arm on the engine, which you do have control over. But be forewarned, all of the roller rocker arm manufacturers use different design geometry, because they never designed from any established standard; consequently, the adjusting screw or pushrod cup is in different angles and heights to the roller tip's location. When you adjust the installed geometry of the rocker arm on the engine, by using different length pushrods or stand heights, you will have different height pushrod locations that will move excessively in one direction or the other - depending on the manufacturer. The lesser of evils is still to have a 90 degree relationship on the valve side of the rocker arm, so that the valve and guide will suffer the least amount of side loading and friction. See: Geometry, Design and Geometry, Installed.  ^

ROCKER HEIGHT: Is the dimension used to set accurate Installed Geometry, by measuring the rocker axis centerline (whether it is a shaft or trunnion), in relation to the roller axis (roller pin) in the CLOSED VALVE position. Whatever the NET VALVE LIFT will be as determined by the rocker's NET RATIO multiplied by CAM LIFT, the ROCKER HEIGHT should be HALF of this NET VALVE LIFT when the valve is CLOSED. As explained in the Installed Geometry page, all measurements must be in reference to a perpendicular angle with the valve centerline, and the valve spring retainer is the best reference plane for this. The first measurement, we've called STACK HEIGHT, accounts for roller tip above the retainer, while the second measurement finds the rocker shaft/trunnion's centerline in relation to this same plane, then added or subtracted as needed in meeting with whatever the NET Valve Lift dictates the MID-LIFT point to be.

NOTE: This is especially easy when setting rocker geometry on engines with hydraulic tappets (roller or conventional), as it allows for accurate measurement of the required pushrod length WITHOUT collapsing the tappet's hydraulic valve from rotating the engine through its valve lift, as done when going directly to MID-LIFT. It is important to be sure the tappet's valve body (and pushrod seat) is "up" as it should be on tappets that have been pumped up from the engine's oil pressure, or from being primed. When using new tappets that have not been primed, you must be careful to set your adjustable pushrod gently to avoid any accidental preloading that would push the plunger down, and give a falsely long pushrod length. The tappet's hydraulic plunger motion is limited to about .050", but not knowing that it is fully extended when checking it this way is enough to lose several degrees of valuable cam timing at the valve.

ROD RATIO: (a.k.a. ROD ANGLE) Refers to the "connecting rods" between the piston and the crankshaft; but specifically, the mathematical division of a crankshaft's STROKE by the connecting rods LENGTH. Ranging anywhere from 1.3:1 (not good) to more than 2.0:1, the effects of rod ratio on an engine's performance and specifically its torque curve are significant. Rod ratio directly affects how the piston speed accelerates from its extreme positions of TDC and BDC (as well as slows down in approaching these points). The PISTON VELOCITY still reaches its same peak value, as dictated by the crankshaft's stroke; but this acceleration toward this is determined by rod ratio, with the smaller values of rod ratio making these quicker, and the higher values slower, respectively. One distinct down side of lower rod ratios, is their inherently higher thrust of the piston's skirt into the cylinder wall. These require more piston skirt clearance compared to a similar stroke engine using a longer connecting rod. Of course, point of fact: if you have two engines of the same model with the same stroke, but different length connecting rods, then of course the difference must be made up with either a different COMPRESSION HEIGHT piston, and/or a greater DECK HEIGHT block. It is commonly accepted that connecting rod ratios of 1.70:1 to 1.90:1 are preferred for most applications. Deviations above and below these values usually have some other negative offset. The higher the rod ratio, the higher the shift in RPMs for peak torque to occur. The more loosely used term of rod angle really has no literal meaning, since no one in engine development actually uses an angular value in describing or deciding what connecting rod lengths are used, not even in their effects of operation.

ROLLER: With roller tip rocker arms, we are referring to the roller which resides atop the valve stem tip. Mythically perceived to actually roll in operation, it does NOT. It merely provides a pivoting point for the direction of operational load imposed by the rocker body, much the same way as a roller lifter provides over a conventional "flat tappet" cam follower. With CAM lingo, we are referring to a ROLLER TAPPET. See: CAM FOLLOWER.

ROLLER PIN: Sometimes referred to as the "roller axle," it is the device which the roller itself pivots around. Various manufacturing methods to attach this axle to the rocker body are used, from pressed-in (interference fit) to "rivet" or "nail head" designs which use a "C-clip" on the other end.  ^

SAE: Standard Automotive Engineers. An automotive standard of engineering used for many aspects of manufacturing dimensions and tolerances beyond the automotive sector. Fine threads follow SAE specifications.

SECONDS (Angular Terms): See: ANGLES.

SELF ALIGNING: Referring to rocker arms, is a rocker arm whereby the alignment of its tip with the valve is assured by protruding rails (on shoe tip designs) or washers (on roller tip designs) which straddle either side of the roller, and have an outside diameter greater than the roller for which they are aligning. The washer (or rails) purpose is to hang to the side of the valve tip and prevent the rocker arm from misaligning during operation. Standard roller tip rocker arms by numerous manufacturers jumped onto this OEM concept by simply juggling the roller width, or slot widths of their original "non-aligning" rocker arm designs to allow for these washers, which usually only extend anywhere from .030", to .050" or .060" along side the valve tip.

NOTE: The "self aligning" rocker arm has ONLY ONE PURPOSE: REDUCE COSTS of MANUFACTURING and ASSEMBLY. There is absolutely NO PERFORMANCE ADVANTAGE to this concept. It is deemed by MEI as a compromise which has risks to its use on any real high performance application. It is too easy for the roller tip to jump off alignment if any over revving, valve float, or other high RPM challenges come up during high performance operation. For this reason, MEI does not make the easy modification that we could in offering this option to our PVS rocker line. We strongly suggest that the addition of GUIDE PLATES be added to any application which had self aligning rockers from their OEM.

SHAFT: Self descriptive, we refer to it here as the roller rocker arm's mounting type, rather than "trunnion," as used in "stud" mounted designs. Shaft mounting consists of one or more rocker arms mounted upon each shaft which is usually secured to an individual stand by two or more mounting bolts, or studs, on either end of the shaft. Shaft systems principle virtue, when designed accordingly, is tying the rocker arms together in forming a more rigid assembly, thus reducing harmonics to the rocker body and valve springs; their principle foe to longevity.

SHAFT SYSTEM: (Rocker lingo) Refers to STAND MOUNTED rocker arms, whereby the first person to coin this phrase was trying to make a distinction between rocker arms that use a removable shaft, from a rocker arm that has a fixed shaft, more aptly referred to as a "trunnion." In all reality, a trunnion (as pertaining to rockers) is a "shaft" and a shaft is a "trunnion." But this slang distinction stuck, and has been repeated for decades. We prefer "STAND SYSTEM" since it is a more accurate definition.  ^

SHIMS: A variety of different strips of metal formed in one of several ways, and of many thicknesses, often custom made from sheet steel or aluminum. They've been around since the beginning of time for adjusting many things on an engine, during engine development; but with regard to rocker arms and valve train, they have been used to change the height of rocker stands on shaft mount (what we call "stand mount") rocker systems. Because every rocker arm manufacturer we know of has always used stands that were too short for MID-LIFT geometry, they have always used this means to elevate their stands up to what each engine builder's requirements were. We consider their use to be band-aid engineering.

SHOE TIP (slang): Refers to the fixed, radius contact pad of the rocker arm which lays upon the valve stem's tip, and directly pushes against it during the operational process. Primarily used on OEM applications for their simplicity.

SHORT SIDE RADIUS: Cylinder head lingo, that refers to the bottom side of the intake or exhaust port, where the bend to (or from) the valve pocket area is made to connect the main port length's passage way, or port "floor."

SHOT PEENING: A manufacturing secondary process whereby select, predetermined sizes of differing alloy "shot" is literally blasted by high pressure across selected surfaces of any ferrous or non-ferrous component, to microscopically hammer down and work harden the surface for increased resistance to surface flaws, fatigue and STRESS RISERS. Although there is no doubt that this process can enhance the surface integrity of some applications; its merits are arguably minor in most cases where adequate materials, manufacturing processes and engineering are properly applied.

SOLIDS: (Slang, cam lingo) Refers to mechanical cam followers, also known as (aka) Solid Tappets or Solid Lifters. See: CAM FOLLOWER.

SOHC: (Head, cam and engine lingo) SINGLE OVER-HEAD CAM. Refers to engines having a single camshaft above the cylinder head, as opposed to traditional engine designs whereby the cam resides in the engine block and uses linkage between it and the head (or heads) for opening the valves. DOHC is DUAL OVER-HEAD CAM and is the same concept except for an engine design whereby the head (or heads) have two camshafts atop the cylinder head, usually dedicated to opening the intake and exhaust valve arrays separately.  ^

SPLAYED VALVE: Cylinder head lingo, referring to valve arrangement with the cylinder head's combustion chamber, as measured in angles along the cylinder head's axis`. Specifically, "splayed" valve array is the leaning of the valves along the "X-Axis" of the cylinder (either end), making the valve stem centerline less than perpendicular with the cam's rotational axis. Rocker stud or pad mounting with this valve arrangement always requires a dedicated, compound angle to achieve accurate arcing trace in line with the valve's motion.

SPRINGS: (Slang) See: VALVE SPRINGS & PRESSURE.

STACK HEIGHT: A term we coined to define the height of the roller pin axis above the valve spring retainer. Used in setting Installed Geometry, it is simply taken by measuring the height of valve tip above (hopefully) or below the top of the retainer, then adding HALF of the rocker's roller diameter in finding the axis. This measurement is then calculated with the second measurement taken from the valve spring retainer top to find the center of the rocker shaft (or trunnion) adding it or subtracting as needed. The final sum of the effort is known as the ROCKER HEIGHT, and this should equal HALF of the NET VALVE LIFT.

STAGGERED VALVE: Cylinder head lingo, referring to the valve arrangement with the cylinder head's combustion chamber, where the intake and exhaust are shifted from each other along the Y-Axis (width) of the head. There is no specific criteria for this definition including or precluding the valve angles to head's deck, but it does mainly apply to valve stem angles which are perpendicular to the cam's centerline. However, compound geometry valve arrays, such as Big Block Chevrolet (where the valve's are "splayed" to the X-Axis) are often included in this term.  ^

STAND: In rocker arms, it is the foundation of what the rocker "shaft" lays upon. In some OEM applications, it is cast into the head and it is more aptly referred to here as a Pedestal. Our description of other companies` "shaft" systems.

STAND OFF: (Induction Systems) The condition of fuel and air that has been pushed back through the carburetor inlets due to REVERSION and inadequate manifold design, to the point where a real and visible fog of this collective symptom lingers atop or at the opening of the induction tract or carburetors. See: PLENUM CHAMBER.

STRESS RISER: A term used to describe the microscopic flaws that develop, usually in a corner, where a concentration of stress and flex will focus, and typically leads to a fracture. Stress risers are reduced or prevented by eliminating sharp corners with use of an increased radii where adjoining surfaces meet (rather than a sharp corner), and sometimes by micro-polishing. SHOT PEANING is another process that is used to "work harden" the surface of components that may suffer from surface fatigue, that can encourage stress risers to begin. Aluminum is especially susceptible to machining flaws that leave these corners where high loads occur, since aluminum is most volatile to FATIGUE.

STUD: This refers to any single bolt type fastening device, usually made of steel in varying grades, sometimes titanium, whereby both ends are threaded. However, in VALVE TRAIN terminology it is the rocker arm mounting technique used with many engine designs, which we've always referred to as "STUD MOUNT." Specifically, the mounting of a rocker arm upon a "fulcrum" which resides on a stud mounted to the cylinder head, providing simple and quick production capabilities; although limited in strength and support for high valve spring loads and RPM use. Principally needed with "canted valve" cylinder head designs, such as the big block Chevrolets, 351 Cleveland and 460 Fords, etc.

NOTE: Rocker arm studs are notoriously TOO SHORT for proper rocker arm geometry on many applications. This is a common  fact, and too many engines are assembled with the set screws run down in the adjuster way too far because there is only four or five threads being used (and sometimes even less). In fact, with lesser valve lifts on milder cam engine applications, the problem is worse since the rocker arm needs to be raised higher on the stud to accommodate placing the mid-lift point higher for the lesser valve lift. (Greater valve lifts move the mid-lift point lower in relation to the valve tip.) Do not be surprised to see that you only have a few threads left when you've elevated the rocker body up to set the installed geometry to mid-lift specs. You need at the very minimum an depth of threads that is equal to the diameter of the stud you are using, and ideally you will want 1.5X that diameter or more. Various companies offer "extended length" studs, but still some are not enough or they are too much.

One last point about ROCKER STUDS; as we all know they come in both 3/8" and 7/16" diameters. Many people switch to the 7/16" for more strength, which is true that it has. But in MOST CASES, especially when correct geometry has been used, 3/8" is plenty adequate if using a strong alloy stud. The DOWN SIDE to 3/8" studs, is they are more likely to be TOO SHORT for correct MID-LIFT geometry, more so than the 7/16".

STUD-BRIDGE™: An innovative Patented device, made of a single bar of aluminum using precision machining to accept specially machined 1/2" Dia. bolts that have half-moon cutouts which overlap extended length, precision ground adjusting nuts for securing the rocker arm mounting and adjustment to an stud mounted, OHV engine design. Follows the same intent of what is known in the trade as a "stud-girdle." (Only better.) STUD-BRIDGE™ is also a pending Trademark (™).

STUD GIRDLE: Is a device made usually of two or more bars of aluminum having half moon cutouts in each bar to provide an elliptical grip upon extended length adjusting nuts when these two bars are tightened together by cross drilled and mounted bolts. Its purpose is to increase rigidity to rocker arm mounting studs on stud mount valve trains, which would otherwise be free floating and flex under high spring pressures and RPMs. It dates back to the 1960's.

SWEEP: A term used in describing the "roll" (or scuffing) across the valve stem tip, or its "sweep across the valve."  ^

SWIRL: Cylinder head lingo, refers to COMBUSTION CHAMBER design for cross flow that induces a preset flow pattern into the combustion chamber which is hopefully controlled into a rotating, high speed "swirl" that fills the chamber in a more complete but aggressive way, which does two things: first, follows the more violent patterns of turbulence created in a wedge combustion chamber, that has better flame propagation tendencies; and second: sweeps across the piston dome during overlap to scavenge the cylinder better. The piston dome, is viewed under these philosophies as the "floor" of the combustion chamber, and its shape too is often critical in the success of this often figure "8" shaped chamber. It made its debut across the aftermarket thinkers of cylinder head design in the mid-seventies, but is still a near black science to many. It does work, but its merits are really beneficial on high-tech, high-efficiency engineering, or aftermarket heads that have pursued this. Most cases of this latter example (aftermarket heads), the combustion chambers are compromised to follow some of these creative shapes, but still fall short of applying them to their fullest potential. Swirl technology depends on both: the intake and exhaust port shapes, as well as their attack angles to and from a specific combustion chamber design, that works with the piston dome's profile.

SYMMETRICAL (ASYMMETRICAL): With rocker arms (or cams) this refers to the equal division of motion, in mirrored fashion. Whatever the first half of the total motion is, is reverse duplicated on the second half; more simply: "balanced," "proportional" or "even." "ASYMMETRICAL" is the opposite of symmetrical; having no balance or symmetry. With Asymmetrical, a term often used in cam designs, the opening cam lobe's acceleration profile is different than its closing side profile. This is done to get what is known as "area under the curve" during the greatest points of breathing capability, and then closing the valve where the most optimum point of ending this breathing is believed to be. Since the balance of these two goals is not symmetrical in proportion to the cylinder head's breathing characteristics, neither is most cams used in today's racing.

TANGENT: Is the term used to describe the actual meeting place of any intersecting lines of shape, form or physics. In rocker arm terms, we use this often to describe the axis points created within the pushrod tip or cup and the imaginary line that stretches between this point and the rocker arm's pivoting trunnion or shaft. Same for the imaginary line between roller's axis and the trunnion. Where these lines intersect is their "tangent" points.  ^

TAPPET: Also "lifter." See: CAM FOLLOWER.

TENSILE: Any force applied to or resultant from loads placed in an opposing direction. (Stretch) Opposite of COMPRESSION. It is also the most reliable form of predictable tension application in securing any component where maximum strength is required. In its ultimate use with fasteners, the actual stretch effect of TORQUE being applied is actually measured in its stretched value, rather than the torque of the applying force, which is less accurate.

THROTTLE BODY: (Induction Systems) Refers to the controlling "butterfly" valve and housing which governs the air/fuel supplied to the engine's intake system tract. More specifically referring to Fuel Injection (FI) systems, where the timed fuel delivery is automatically controlled, in most cases by a computer and prescribed size injector orifices located in either the throttle body itself (not as efficient), the intake manifold or the cylinder head; which are synchronized with the degree of butterfly opening created within the throttle body that is linked to and controlled by the driver's "foot." In these latter examples of fuel delivery, the throttle body is merely controlling the air into the system. In all cases it is designed with an ultimate limited capacity based upon its inside diameter that equals the butterfly's diameter.

TOP DEAD CENTER (TDC): Refers to the position of the PISTON (and crankshaft) in relation to the stroke of the crank. It is the ultimate top of the crankshaft's stroke, before initiating its travel back down.

TORSIONAL TWIST: Is the rotational flexing that occurs around an axis of any linear object where a force of torque has been imposed upon one end that is greater than the absorbed torque transferred through to the restrained or driven opposite end of said object. Camshafts experience this, so do crankshafts (to a lesser and different degree). With camshafts however, the effect is an actual retarding of timing between cam lobes at the opposite end of the cam from those near the driven end. Engine builders for many years have experimented with grinding the cam to anticipate this, by place the rear cylinder cam lobes 2, 3 or more degrees advanced from the lead cam lobes near the timing end of the cam, so when the engine is accelerated to peak operating RPMs where the torsional twist is at it's highest, the now retarded rear cam lobes have been evened up with the lead cylinder lobes.

TORQUE: A force applied to or from a rotating axis; measured at a fixed distance from that axis. See: HORSEPOWER.

TRACKING VELOCITY: Is the accuracy of how well the rocker arm converts linear information into radial motion from the cam. The closer to "in-line" that tangents remain with the linear path of the pushrod, the less deviation occurs from the cam's acceleration dynamics. Tracking velocity can be altered by placing the 90 degree tangent either closer to open valve or closed valve, with reverse consequences to each; but doing either alters the cam's profile at the valve, diluting or accelerating the valve's response of the cam information.

TRUNNION: This refers to the singular steel "shaft" within the rocker arm of a "stud" mounted design, having 2 bearings, 1 on each end, with a hole in its middle to accommodate the cylinder head's rocker mounting stud. Because a trunnion is also a SHAFT, we choose to call the other rocker arm design "STAND SYSTEMS" rather than "shaft systems." (If you're going to try to be known for accuracy, might as well go all the way.)   ^

TURBINE INLET TEMPERATURE (TIT): (Turbocharger) The critical temperature to know on turbocharged engines seeking maximum performance with minimal risk of overheating the Turbocharger, particularly on engines whereby the fuel to air ratio is adjustable. Measured by a temperature probe usually mounted within (but not into) two inches of the turbocharger's Hot Section inlet, where the collective primary tubes of the engine come together, it is a commonly accepted operating formula for the "sustained" maximum temperature to never exceed 1,650° Fahrenheit, and in most conservative circles of use, this limit is held to: 1,550° Fahrenheit. It should noted that these are collective flame/gas temperatures within the exhaust system, not component temperatures of the system, which run several hundred degrees less. However, the cast iron hot section of turbochargers operating in these upper temperature limits will literally become a glowing orange/red, exerting immense radiant heat to all components around them, especially in confined engine bay areas, where ventilation, space and heat shielding have not been implemented as part of the installation. Another important, but less noted (and used) term refers to temperatures taken on the discharge side of the Turbine section, known as TOT (below).

NOTE: These temperature limits should not be confused with Diesel engine designs, which operate much lower, nor should it be confused with Exhaust Gas Temperature (EGT), which will usually be held to no more than 1,450° Fahrenheit, and measured by carefully mounted probes that are approximately 4" out from the cylinder head's exhaust port flange to avoid direct contact with flame front core. DO NOT CONFUSE EGT with Turbine Outlet Temperature (TOT).

TURBINE OUTLET TEMPERATURE (TOT): (Turbocharger) This is the exhaust gas temperature exiting the Turbocharger after driving the Turbine, which is then passed through a connecting exhaust pipe and muffler on engines so equipped. As with the TIT, the temperature probe is mounted within 2" (two inches) of the Turbocharger's Hot Section outlet. This probe location, which sees the Turbo's cumulative temperature from operation under all types of loads, is instrumental in determining cool down time before shutdown of the engine; which is often required with most applications that have been operating at a high load, in preventing premature wear and failure. This cool down time can be anywhere from one to five minutes (or more) of low or no load operation before killing engine. This brief time is often determined by various manufacturers of so equipped vehicles, and it is all time required to lower the TOT temperature (which is sometimes called the EGT, because it may be the only temperature probe on the engine's exhaust) by 250° to 350° Fahrenheit, making all the difference in turbo life when it has been taken to these higher extremes under load.

TURBINE WHEEL: (Turbocharger) The carefully designed impeller of the Turbo's Hot Section which is the driving first stage of the Turbocharger's operation. The Turbine Wheel, which is driven at super-high RPMs of up to and over 200,000, often operates from sustained exhaust gas temperatures ranging between 1,400° to 1,650° Fahrenheit, but rarely more without incurring imminent damage. This temperature is measured as TIT, for "Turbine Inlet Temperature," whereby a probe is usually mounted within two inches of the Turbocharger's hot section inlet.

TURBOCHARGER (aka: "Turbo"): In its simplest definition, turbo-charging is the process of forcing air into the engine's induction system through use of the engine's exhaust pressure. Turbochargers are essentially two carefully designed circular air chambers that work together through combining a common shaft which has two impellers, one on either end, but each with specific blade geometry for the two extremely different tasks they do. One half of the Turbocharger is a cast iron Hot Section that connects directly to the exhaust primary tube collector, or cylinder head's exhaust manifold, and receives high pressure, high temperature exhaust gas pulses from the engine's exhaust. These gases act upon a Turbine Wheel (above) to drive it at very high speeds, through a common shaft which then drives a Compressor Wheel made of impeller blades formed to a  very specific, complex geometry with two stages of operation. The first stage is dictated by a smaller (minor) diameter group of blades known as the Inducer, which grab and pull the incoming air drawn from the air cleaner (or an ambient intake source), which then "compresses" the air, which raises its temperature by 100° or more, and forces it out through a second stage of blades with a larger (major) diameter, known as the Exducer, that force the pressurized air through a precisely shaped, spiraling passage in the aluminum "compressor" housing, also referred to as the Cool Section. The compressed air is [usually] then sent through an aluminum "heat exchanger" known as an Aftercooler, (but often incorrectly referred to as an "intercooler") where the air temperature is reduced before being sent on to the engine's induction and throttle body. Both sections, the hot and the cool, have inlets and outlets; and in addition, both have a means to control the pressure development of both sections. On the Compressor side (Induction), a Bypass Valve is used, also known as an Over-boost Valve, which allows excessive pressure in the intake tract to escape when a predetermined value has been reached. On the Turbine side (Exhaust), a bypass section designed integrally with the cast iron housing, or separately connected to the discharge side of the hot section, known as a Wastegate, provides excessive pressure from the engine's exhaust system from running the Turbine faster than necessary for the loads which it is designed for. Literally, the inlet side of the Turbine is connected through this alternate passage to the outlet side, and thus "bypasses" the Turbine itself, dumping directly into the  final exhaust system. Some wastegates are adjustable, either manually, or automatically, to compensate for different temperature and/or atmospheric pressure conditions.  ^

TURBULENCE: Cylinder head lingo, referring to air flow patterns and dynamics. Specifically, it is the point at which LAMINAR airflow patterns leave controlled, smooth layer (or swirl) flow characteristics and collide in uncontrolled, random (and violent) shifting patterns, capable of motion in all directions, including "up-stream." Some poorly shaped ports inhibit these bad dynamics throughout the majority of their valve lift ranges. Two of the largest culprits to such ports are PORT WINDOWS that are too straight across the SHORT SIDE RADIUS, creating a "ski ramp" effect, where airflow leaves the port floor as it approaches the valve bowl area, effectively having no flow across the near side perimeter of the valve head because air is literally skipping over it. The second is too big of a transition from a smaller CROSS-SECTIONAL-AREA upstream; where the flow slows down too quickly as it enters this larger area; gets in its own way and tumbles.

TWISTED ROCKERS: This is a term that we use to describe an abnormal operating condition of a rocker arm being turned, or "twisted" on its mounting axis from the correct alignment required to maintain its reciprocating arc true and in-line with the valve's centerline. This is most prevalent on STUD mounted rocker arms where the pushrod has been offset and the stud location has been shifted to the same direction, an incorrect flaw done too often by cylinder head manufacturers, forcing the rocker arm to now be rotated out of alignment with the cam's centerline, as is needed on in-line valve array cylinder heads that use roller tip rocker arms, where this error most often occurs. NOTE: On OEM style BALL Fulcrum mounted rocker arms this problem is nearly non-existent, because with a ball fulcrum, the rocker arm can shift its twist sideways to keep the rocker's valve tip end flush atop the valve, even though it is twisted on the stud. The ball allows it to self align on two axis, where the needle bearings of a trunnion design roller rocker can only control one axis; and their axis must ALWAYS be in parallel alignment with the cam centerline on in-line valve array engines. Period.

UNDER-ARCING: A term we use to describe the radial path a rocker arm makes on either end which reaches a 90° relationship to its axis with the valve centerline BEFORE reaching the MID-LIFT point of total sweep; causing it to cut beneath itself as it approaches peak lift. See: OVER-ARCING (above).

VALVE BOWL: (Cylinder Heads) Refers to the immediate material making up the inside diameter of the cylinder head port, just beneath the lowest angle of a valve seat, where the port is traditionally its smallest cross sectional area on a performance head; although many OEM heads, or stock type aftermarket replacement heads may in fact have even smaller ports further upstream (on the intake), to keep velocity high in near stock performance application engines. This area of space that is called the valve bowl is about 1/4" to 3/8" below the valve seat, depending on the head model. It is a very critical standard that works in harmony with the valve diameter to establish THE STANDARD for which a cylinder head's ultimate potential and efficiency are established; as well as the yard stick for port dimension limits that should not be exceeded in pursuit of high efficiency laminar flow ports, on any racing engine. The valve bowl and the valve are GROUND ZERO for all port dimensions and limits. See: CROSS-SECTIONAL-AREA for references.

VALVE CENTERLINE: In referring to rocker arm geometry, the valve's centerline is the most important angle to reference from. This is what must move with the least amount of side load, friction, or wasted motion. Additionally, it is the bearer of the valve spring. And when the various levels of motion, heat and impact are added together; the last thing that is needed - is to ask the valve to straighten itself against an over-arcing rocker arm that is throwing its side loads down at 80 cycles per second. The VALVE MUST move STRAIGHT up and down in its guide, and nothing else. The valve centerline is the basis for assuring the rocker arm operates accurately.  ^

VALVE CLEARANCE: Is a term that refers to PISTON to VALVE (PV). A critical dimension, that is varied in some regards due to the engine application being built for (i.e., Drag Racing, Circle Track, Street, etc.), as well as the RPMs, camshaft type (Roller Tappet, Flat Tappet, Hydraulic Tappet), and also on the engine builder's preference. Some engine builders like more clearance than others, some don't have many options for how much clearance they can build for, due to cylinder head, compression ratio, valve lift and a few other factors that are more important than a certain safety value that exceeds what they may be confident they can "get by with." But it is a dangerous game. The published "norm" for many years has been .080" for the Intake valve and .100" or more for the Exhaust valve. The BOSS 429 Ford Hemi engine requires even more, by as much as .020" on each; while some high RPM small block engines using Titanium valves and reasonably high, but not insane valve lifts can get by with as little as .060" and .080", respectively. If you have the luxury of dictating what you need, without losing compression ratio from over cutting the pistons to accommodate this, then by all means run the higher figures. If you accidentally over-rev the engine, then this extra margin will have paid for itself.

NOTE: Piston to Valve clearance occurs during the OVERLAP cycle, when the piston is rising and chasing the Exhaust Valve UP as it is closing, while at the same time the INTAKE Valve is opening and coming down to meet the rising piston. Piston to Valve clearance is always affected by advancing or retarding the cam, and you can gain or lose accordingly. When changing to MID-LIFT rocker arms, you will also see more "area-under-the-curve" (quicker valve response at less crank position), so it is easy to lose Piston to Valve clearance here, if you're already at the minimum. Another key point: Do NOT make a parallel between increased VALVE LIFT and LOST Piston to Valve clearance. Because PV clearance happens at Overlap, it is not uncommon to see more valve lift have a minimal effect on valve clearance, anymore than it is not uncommon to LOSE Piston to Valve clearance changing to a cam with LESS valve lift. It all revolves around "timing" - and the PEAK valve lift often have little effect in this regard to PV clearance.  ^

VALVE FACE: (Cylinder Heads) Refers to the precision ground surface of the valve that actually seals against the mating surface of a like angle ground into the cylinder head, known as the VALVE SEAT. Usually, there is a 1/2 degree "positive" angle of interference between these two mating surfaces, meaning that the valve's outer edge contacts the seat first, to allow for valve head flex and efficiency of the valve's perimeter breathing (maximum diameter being used).

VALVE HEAD: The actual major diameter of the valve which encompasses the valve face angles used to seal against the valve seat of the cylinder head.

VALVE LASH: Is the mechanical clearance of free play between the valve tip and the rocker arm, designed to provide adequate clearance in keeping the valve's closing un-interfered with from operational heat and dynamics. Traditionally applied in meaning to mechanical cams, such as solid roller and solid flat tappet designs, it can technically be applied to the hydraulic cam followers too, even though this lash is controlled by hydraulic valve compression within the tappet, through metered bleed-off of the engine's oil pressure. For purposes of continuity across cam applications, we always use ZERO LASH in calculating ROCKER RATIOS.

NOTE: Our LASH RECOMMENDATIONS are "tight." Always have been, long before it became reduced to the lash recommendations of some cam companies. I would like to say that we pioneered it, but we didn't. My first introduction to tight lash, was in 1970 with the BOSS 429, after falling victim to the published rhetoric of the factory's 1960's lash thinking of .020" to .025" on the Ford 427 FE, from two years prior. The lash settings for the BOSS 429 recommended .013" intake and .015" exhaust. This could have been argued against by older logic, noting that aluminum heads "expand more," so they need more lash. Motorcycles, my other love, routinely set lash by rotating the pushrod freely! That's tight! Therefore, we recommend valve lash settings for the INTAKE: .012" to .015", and the EXHAUST: .015" to .018" for ANY mechanical cam. Especially if the cam companies recommend more! The tighter lash softens shock, adds a few degrees of duration (which most people can use, because cam companies routinely cut this short); and lastly will require setting less often, if you've done it correctly. If you need more lash, you've got the wrong cam.  ^

VALVE LASH, SETTING (Tech-Tip): For decades good engine builders have had to blame other things while trying to figure out why their lash settings wouldn't stay locked down. Either they break the set screws in trying to torque them down too tightly, or the lash won't stay set because they weren't set tight enough. Most 7/16-20 high tensile fasteners can easily withstand 35 to 40+ lbs of torque, providing they are tightened down properly. Rocker arms using a lock nut on an adjusting screw (STAND MOUNT SYSTEMS), or rockers that have adjusters with an Allen set screw (STUD MOUNTED), can both benefit from a little known technique we call PRECEDING, which will assure they stay locked down, and avoid breaking the set screw in the process:

FIRST: With the set screw loose, use a boxed end wrench on the adjusting nut for stud mount rockers (or the Allen wrench used on the adjusting screw of the STAND mounted rockers), to set the valve lash as you normally would, being careful to take note of where the wrench points when you have the setting you want, as if it was an hour hand on a clock.

SECOND: Now rotate the same wrench counter-clockwise approximately 1/4 or 1/3 of a turn and stop. Then turn the SET SCREW CLOCKWISE to take up the free play, until it bottoms out (or the outer lock nut if STAND Mount rockers), remembering that your setting position is 1/4 or 1/3 of a turn loose.

THIRD: Lastly, now use BOTH hands to hold the Allen wrench and the boxed end wrench to turn BOTH the set screw and the adjusting nut (or adjusting screw and outer nut) simultaneously, tightening both to the same hour position your boxed end wrench was at when lash was set as you desired before locking down. In this final step, you are applying the locking torque in a way that tightens both the adjuster and the set screw (or the outer lock nut and adjusting screw if it's a STAND Mount system) so they preload within themselves EVENLY, without forcing unnecessary pressure to either one individually.

NOTE: The reason for the varying preload in the second step, where you back off between 1/4 and 1/3, is because all mechanics have a different "feel," and you will need to experiment to precisely stop on the exact spot which preloads to the desired lash setting you measured when there was no preload in the first step. After two or three tries, you'll get it and the remaining rockers will set quick and easy. Good luck. --jM

VALVE LENGTH: (Cylinder Heads) Is an easy to understand, self descriptive term with a couple of notes that should be kept in mind, especially for novice engine builders. First of all, in buying aftermarket valves, these are often referred to by an extended length over stock. So you'll often hear "hundred long," or " two hundred long" meaning .100" or .200" longer than stock, respectively. But be careful, because on some heads models of some manufacturers, these "stock lengths" changed over the years; so this can be a moving target if you're not careful. It's better to have a way of measuring what you really need, rather than just fitting pieces together because they are the most commonly used, or in stock. Too many times, engine building comes down to compromises on doing it right, because you're faced with buying parts that had a domino effect which forced you to buy related parts that weren't ideal. Most of the better valve manufacturers easily provide the OAL (Over All Length) for their various part numbers so you pick and choose with more reference to what you really need. If you don't find this information in their catalog, then just call their technical staff and ask. If they don't know, find another manufacturer! Secondly, all valves are measured and sold by this OAL, which includes the VALVE MARGIN; and this may be a different spec between some valve models and/or manufacturers. So be sure to ask about this when buying, and make sure you don't lose .020" from specific height you are looking for that is measured from the valve seat - which is what really matters. With the OEM of cylinder heads (Ford, Chevy, Dodge, etc.), this is known as the GAUGE HEIGHT for the valve seat.

NOTE: Choosing the right valve length is very important to rocker arm geometry; especially our MILLER MID-LIFT ROCKER ARMS. We established more than 25 years ago standards of lengths for various cylinder head models based from a height of the VALVE TIP above the SPRING PAD -- so we could place the roller path as close as possible to middle of the valve during its arc. The spring pad is often a taboo surface to mess with, and always the foundation for determining valve spring heights, valve spring retainer designs and their ultimate valve spring installed height, which is necessary for adequate coil bind clearance. So it is only logical to use it for designing specific valve tip heights, over any other precedents. In addition to this however, regardless of which rocker brand or style you use, the valve tip is critical as the starting point of leverage in setting its contact angle for the rocker arm. Some rocker brands (actually all of them, except ours - we'll modestly note), have their stand heights and rocker body designs engineered with no relative choice of where the valve tip is at. They require that engine builder just figure it out for themselves; and unfortunately, valve lengths are just ordered randomly, too long, and the "SHIMS" (which we hate to see used) are always excessive - especially after the engine builder has really figured out how to set their installed rocker geometry.

VALVE LIFT: Is self descriptive, as the amount of lift the valve operates with. What should be noted though, is that this term is used (and unfortunately trusted as accurate) on CAM CARDS or in cam sales, when in fact the reality of it is that it is often NOT accurate. These are more appropriately termed "theoretical" Valve Lift, while an engine's true valve lift is called "NET" Valve Lift. Net Valve Lift is usually LESS than "advertised," but sometimes when improper rocker arms are chosen or incorrectly installed, can be MORE than expected (although rare). In either event, expensive errors in critical clearances, like Piston to Valve can result from "trusting" theoretical calculations. Always check. Valve lift is the result of a multiplied RATIO by the rocker arm of CAM LIFT. Never really standardized before MID-LIFT, "NET" valve lift at "ZERO LASH" is the best ground zero method to calculate net valve lift by, because it leaves the variables such as "lash" up to the engine builder in determining the engine's final valve lift specifications.  ^

NOTE: For CHOOSING VALVE LIFT, see: CYLINDER HEADS to get the whole perspective.

VALVE LIFT DYNAMICS (VLD): This is a term we use to describe the actual, net, valve lift events, in the same way they are analyzed at the cam: LIFT, DURATION and VELOCITY. Rate of acceleration and deceleration is the more accurate description of "velocity" as we use it here. What is so critical about this perspective, even more so than the cam's events, is this takes the entire valve train system into account, and gives us the magnified results of cam information as it is translated through the rocker arm. Any changes in rocker arm geometry, either design geometry or installed geometry, directly and significantly influence the VLD, and the TRACKING VELOCITY. It is through plotting the VLD on graph paper that AREA-UNDER-THE-CURVE is seen, quantified and analyzed for what is really happening to study HORSEPOWER data.

NOTE: Tracing the valve motion directly to establish a given engine's VLD is crucial to properly understanding the end consequences of cam designs and rocker geometry. CHARTING this with graph paper, which illustrates the Area-Under-The-Curve clearly is necessary. There are TWO methods for doing this: (1) Use an even increment of CRANK degrees (horizontally on the graph to mark TIME) and write down the amount of VALVE LIFT for each increment you've chosen; such as each 5° of rotation; or (2) Use an even increment of VALVE LIFT, such as .005", and write down the degrees of crank rotation seen. Both illustrate the same results, but from opposite perspectives. The second method will clearly show gains and losses of crank rotation due to over-arcing rocker geometry, and it is a test that can be done on any engine, with any rocker system, by merely doing it twice and changing the pushrod lengths a reasonable amount between the two tests, such as .050" or .100". Following that, you won't have too much trouble understanding how important precise DESIGN GEOMETRY rocker arms become, as well as exact length pushrods for proper INSTALLED GEOMETRY.  ^

VALVE LOCKS: See: VALVE SPRING RETAINER.

VALVE MARGIN: Is the width of material on the valve head's major diameter, which lays between the valve's bottom face and the valve seat face. It is very critical to valve longevity, to disperse (or absorb) heat in conducting it to the cylinder head's valve seat. It is also critical to air flow, mostly at low lifts, and especially the exhaust, but also has an effect at mid-range lifts, depending on the port, combustion chamber and application. It is sometimes used in conjunction with a radius on the valve's bottom (chamber side) to induce flow across the exhaust, but most testing has shown this not to be beneficial, and in fact opposite. The argument will continue. One thing that few people argue are the minimum widths before throwing the valves away, or re-cutting their outside diameter to recover a too thin dimension, and these specs are: INTAKE. .030",  EXHAUST .050". Although some engine builders have got away with .020" INT. and .040" EXH., anything less is real risk of burning, and anything in these ranges has already had a detrimental effect on air flow across the seat. We never rebuilt a head that was less than .030" INT. and .050" EXH. New design specs were always .050" INT. and .085-.100" EXH. Just my personal preferences. (See: VALVE TIP LENGTH for illustration of margin's location.)

VALVE POCKET: (Cylinder Heads) Refers to the area just beneath the valve seat area of the cylinder head. It is usually 10% less area in size than the valve itself is, on HEMI style heads, and 12-15% less for WEDGE heads. It is also the transition portion of the port's short and long side radius` to the valve seat's bottom angles.

VALVE SEAT: (Cylinder Heads) Refers to the precision ground angle in the cylinder head that provides a seating surface for the valve's VALVE FACE to close upon. It is usually .040-.080" wide for INTAKES, and .080-.120" for EXHAUST valves. Different engine designs, applications, materials and "tricks" by various engine builders will vary these general dimensions in both directions.  ^

VALVE SEAT INSERT: (Cylinder Heads) Refers to the independent component which is separately installed into a cylinder head in providing the precision valve seating surface, necessary with all non-ferrous (aluminum or magnesium) cylinder heads, or any iron heads which required an alternate valve seating material of higher strength, higher operating temperatures or simply restoration from original valve seat damage. Valve seat inserts are made of a variety of materials, from low tech iron to high tech beryllium, stellite, and other high temperature alloys. With titanium valves, these extra hard nickel and cobalt alloys can't be used without beating the valve faces to death. Emphasizing the importance of choosing the proper material in working with different fuels and various alloy valves.

NOTE: Valve seat inserts must be installed carefully. Typically agreed in the industry that with aluminum heads, an interference fit (press fit) of .007" to .008" between the valve seat insert's outside diameter and the head's valve seat insert "pocket's" inside diameter. The machined surface of both must be concentric within .001" and smooth. Lastly, the best way to install them is to "freeze shrink" the inserts with dry ice and alcohol, or liquid nitrogen (harder to come by and more dangerous to use around clumsy hands). TRICK of the trade: REMOVING valve seat inserts is easy. Using a TIG welding torch to heat them up by circling around them about three or four times will anneal them; making their press fit dissolve and they will literally fall out of the head. Be careful not to keep the TIG tip close to the insert and avoid excessive heat going into the head which can hurt its heat treat. Yes, aluminum heads have a heat treat.

VALVE SPRINGS & PRESSURE: (Cylinder Heads) Valve springs and their pressure are notoriously misused. There are two interesting things to keep in mind about valve springs. One: They are a torsional component; meaning...when they open what is really happening is that their individual wire is "twisting" on its own, coiled axis, as the spring is collapsing. The second point of interest, is that valve spring pressures, after a point are not really increased because of MASS, INERTIA and RPMs accelerating these things. Valve spring pressures are increased to overcome the extreme waste of valve spring pressure WHEN HARMONICS reach their critical state. Harmonics come in a series of waves which vary in degree of their intensity. With the first waves of harmonic influence, usually around 7,300 to 7,600 rpm, the end coils of some springs begin to suspend themselves in space. The end coils really don't work anyway, they just support the coils in between. But they begin to leave their mating faces (head and retainer) before anything else starts going to hell. When the second level of waves occur, usually 500 to 1,000 rpm over the first, the spring now has about 1/2 of its coils actually compressing anything, while the rest are dancing all over. The increased spring pressure compensates (in a bad way) for this loss since the end coils are no longer seated in place and the middle coils are fighting each other. Parts are dancing all over the place. Retainers slap around with their locks, spring shims and cups start dancing off their pads and rotating. It's a mess. Harmonics are a direct consequence of many factors, which are exaggerated by BAD ROCKER GEOMETRY. But one "constant" element of harmonics, which makes them somewhat predictable, is the metal alloy springs are made from. They have a frequency which is different in steel, than titanium. Titanium's ultimate frequency is much higher than steel, ANY steel. The ultimate valve springs are titanium, solely because of their molecular influence with frequency. In fact, I know of several tests that have proven as little as 85 lbs of TITANIUM spring pressure can control RPMs with no valve float, beyond 9,300 rpm on a radical cam small block Chevy. The same engine, with steel springs, went into harmonics and valve float just over 8,500 rpm, and needed over 225 lbs of seat pressure to do this. I make this point to remind engine builders that HYDRAULIC roller cams don't need 150 or 160 lbs, or a flat tappet cam doesn't need more than 135 to 140 lbs of seat pressure, on valve trains that have "proper geometry" and "stable" cam profiles. Meaning acceleration rates, slightly beneath the Space Shuttle upon entering the Earth's orbit. It takes HORSEPOWER to turn over extra spring pressure. Use only the spring pressure you need, to do the job. Keep non-mechanical roller tappet profiles to not more than 140 lbs of seat pressure, and not more than 385 lbs over the nose. Mechanical rollers are another animal. Not necessary in most cases of competition, even though the cam companies love to sell them to you. One last thing: See: COIL BIND.  ^

NOTE: There's been a trend by cam companies to misinform engine builders about SPRING PRESSURE on FLAT tappet profiles that use high tech solid (mechanical) tappets they need for various cam profiles. Rather than naming names, pointing fingers, let me explain a couple of things that makes this BS.

#1: There is a specific known value of cam/tappet acceleration that can't be exceeded, because of the diameter of the tappet. THIS is the limiting factor on a flat tappet came. Unless "mushroom" tappets are used, this limit won't change. How quickly the cam accelerates is what determines "area-under-the-curve" which is how fast the valve will open to s greater value for each degree of crank rotation. The inertia created by the acceleration rates which are limited to the .842" Chevy and .875" Ford tappet diameters is below, as it has been for 50 years, the values needed to justify 175 lbs (and more) of seat pressure springs.

 #2: Increased Valve Spring Pressure is used to COMPENSATE for the LOST spring pressure that occurs from HIGH RPM HARMONICS, which are ABOVE 7,800 RPM, and has little to do with substantially controlling valve float on engines running below 7,800 RPM.

#3: The cam companies that are recommending crazy valve spring rates on flat tappet cams don't have a clue about the value of correct rocker geometry. Therefore, all their data is meaningless. Whatever they say a cam needs for valve spring pressure was determined (if at all) from combinations that had over-arcing rocker geometry which consumed vast amounts of energy going to the wrong places because the rocker arm wasn't doing its job efficiently. Period. Secondly, the cam company making such recommendations is also lacking the data needed on your cylinder heads to know if they need, or can use a fast ramp cam. 90% of the heads sold by the top manufacturers run just fine, and can't take advantage of a truly fast ramp cam that comes close to the "aggressive" version designed in roller profiles; let alone flat tappet profiles.

130 lbs to 145 lbs of seat pressure is all that you need on a well designed flat tappet cam, with correctly installed MID-LIFT rocker geometry (for the least amount of over-arcing and binding). Increasing Valve Spring Pressure to 175 or more pounds is throwing horsepower away, adding unnecessary friction, flexing parts more than necessary and heading your engine for destruction. Are there flat tappet cams out there that require 165 or 170 lbs of pressure to get the most out of the engine? YES. But these are very limited examples where the total combination has been tweaked to the last percent, and RPMS are going to exceed 8,300 or more! If you are NOT going to run 8,500 RPM -- and ABOVE, then considering any flat tappet cam from any company telling you that you need 175, or 200 lbs spring pressure, is a company you should be run away from as fast as you can hang up the phone!.  ^

VALVE SPRING RETAINER: (Cylinder Heads) Nearly self descriptive, the valve spring retainer is a flat appearing round disc-like item that sits atop the valve spring and affixes itself to the valve stem by two, tapered, steel cylindrical halves, known as VALVE LOCKS or KEEPERS. As we all know, valve locks act as a door stop in that their two halves are compressed against the valve stem by a corresponding tapered hole in the center of the retainer, of which the upward pressure imparted by the valve spring forces the tapers to "grip" against a matching male/female groove between the keepers and a valve stem. Valve spring retainers from the OEM are made in steel, but aftermarket uses aluminum and titanium, and are substituted from OEM with any new valve spring package to fit larger springs and often to reduce weight. Two basic angles of LOCKS are most often used, 7° and 10°, the latter believed by many to better resist pulling through on radical cams and high spring pressures. With this rudimentary definition out of the way, we can now make the point of FIT needed in a stable valve spring package.

FIT between the valve spring RETAINER and the VALVE SPRING is critical. Valve springs dance all over the place in an operating engine at high RPM's. Keeping this dancing to a minimum is critical. Keeping their friction at a minimum is critical too. Yet, they are often sold and promoted with tight fitting inner and outer springs as if this is a virtue. It isn't. Valve springs are horsepower robbing heat generators. The "step" that is machined in the valve spring retainer needs to have a near "snap fit" where the spring has no free-play to dance around. Because of the many retainer and valve spring variations and combinations used to fit a wide variety of cam kits, often, compromises are made in matching retainers with spring packages. It is difficult to get the proper, tight fit needed between the steps of the valve spring retainer for ALL combinations. On DUAL SPRING and TRIPLE SPRING combinations, it isn't uncommon to find the outer spring fit the machined step, while the inner springs do not. Or worse, the opposite might be found: the inner spring may be held intact by the retainer's protruding valve lock extension, while the middle and outer springs are loose fitting. STEP HEIGHT is also important for each of the corresponding springs within the outer spring, as each of these will COIL BIND at a different height; and these must be checked individually with the retainer, then as an assembly; noting which coil binds first. Ideally, we liked to see the OUTER spring COIL BIND a few thousandths before the inners. But the step height of the pad where each of the inner springs sits against the valve spring retainer should be of adequate height to preload the corresponding spring when assembled to the outer spring's needed height, and NOT be the first to coil bind in doing so.

VALVE STEM: The longest diameter portion of the valve which rides within the cylinder head's valve guide and adjoins the valve tip with the valve head, also encompassing the keeper groove or grooves, to affix the valve spring and retainer by split locks. This is also a wearing surface through its continued operation, showing signs of this wear along its outside surface as "scuffing" or galling - in worse cases - from a variety of usage reasons. In valve train efficiency, the lack of proper rocker arm geometry becomes evident on the valve stems and valve tips, first, before most other symptoms are seen.

VALVE TIP: Upper end opposite the valve HEAD, where direct contact from the rocker arm is made to open and control closing of the valve on OHV engines. The VALVE TIP is everything; EVERYTHING. It is from here that all true consequences of the rocker arm's geometry occurs; yet, engine builders indiscriminately add longer and longer valves, thinking all they need to do is to use longer pushrods to compensate. (See: VALVE TIP LENGTH for illustration.)

VALVE TIP ADJUSTMENT: With higher end STAND MOUNTED ROCKERS ("shaft systems") using a common shaft for both the Intake and the Exhaust rockers of a common cylinder or an entire bank of rockers for one head, the VALVE TIP is the only way and the BEST way to accurately adjust the rocker arm's pivot point in relation to the valve tip. Since Exhaust valve lifts will differ from Intake valve lifts more often than not, there is no easy way to adjust the rocker's shaft height of one and not the other, since they are joined. Yet they NEED to be set accurately. By varying the VALVE TIP LENGTHS exclusively for this purpose, you establish precise Installed Geometry without compromise. For example: Most Exhaust valve lifts will be less than the Intake, so for an engine with .750" Net Intake Valve Lift, and .700" Net Exhaust Valve Lift, the MID-LIFT point for the Exhaust Valve will be .025" higher (half the difference) than the Intake Valve's MID-LIFT point. So to RAISE the ROCKER SHAFT in relation to the VALVE TIP, without raising the shaft, your only option is to LOWER the Exhaust Valve Tip by .025" in comparison to the Intake Valve Tip. If you need a MINIMUM VALVE TIP LENGTH (as most engines do) for Valve Spring Retainer clearance, rocker body clearance or whatever, then whatever this needs to be must fit the shorter height valve tip of the two valves; in our example's case this would be the Exhaust Valve. So you would add the required surplus to the INTAKE VALVE. From this baseline, you would now adjust your STAND HEIGHT for exactly what is needed to set your shaft's position with the Intake Valve. Naturally on stand mounted systems that have independent shafts and stands per rocker arm, this has no meaning.

VALVE TIP LENGTH: refers to the dimension taken from the end of the valve (tip) to the nearest edge of the KEEPER groove. Usually either .250" or .310", this is critical for standardizing to the retainers and valve springs you choose.

VALVE TIP HEIGHT: This is a term we established long, long ago (in our cylinder head prepping days) to standardized the HEIGHT of the valve tip over the valve spring PAD. It is the best way we know of to keep valve length requirements consistent in designing rocker arms. This is necessary because all heads have either an incline TO or FROM the valve of the rocker arm mounting pad (or studs). Making any deviations in valve lengths impossible to establish an accurate ROCKER LENGTH that keeps the roller in the middle of the valve tip. (See: VALVE TIP LENGTH for illustration.)  ^

VALVE TRAIN: is the most temperamental and demanding category of an engine's design. Most engine failures are the result of something breaking in the valve train, usually a valve spring or rocker arm. But most improvements in engine performance come from valve train development. The most precisely designed and driving component of the valve train, which is aptly referred to as the "heart" of the engine, is precision ground to tolerances less than 2/10 thousandths of an inch (.0002"), and computer designed to tolerances 10 times greater. This precision component works in harmony with three other components of the valve train to open the valve, and determines all of the critical timing for the engine's ability to breathe the greatest amount of airflow, while venting wasted gases, all in an effort to produce maximum horsepower. It is known as the CAMSHAFT.

VARIABLE RATIO (aka: ACCELERATED RATIO): This is a confusing, cam company "hype" term which may be good for selling the virtues of over-arcing rocker arms to unsuspecting engine builders, but it has no bearing in fact; mainly because there is no virtue to a rocker arm exceeding its arc more than necessary. But this "variable ratio" and "quick lift" term is used specifically by some companies (one in particular) to promote a rocker arm which has a different (higher) ratio in the closed valve position, then decreasing in its ratio as it approaches full lift position.

NOTE: ALL rocker arms, technically, have a variable ratio. Rocker arms go through an exchange of virtual "contact" point lengths of their center-to-center dimensions during operation, because the lengths from the axis on both sides of the rocker arm are relative to a straight line dimension which is measured through its axis at a right angle to each: the valve and pushrod. Since these rocker arm ends are rotating around an axis, when measuring them relative to the in-line motions of the valve and pushrod, their contact lengths actually change with this arc; even though the actual center-to-center of these tangent points within the rocker body are fixed. Technically, this varies their ratios a "small" amount throughout their lift cycle, but only when MID-LIFT geometry is utilized. MID-LIFT minimizes this to a predefined, predictable standard which guarantees the ultimate minimum in-and-out motion on BOTH sides of the rocker arm. If different geometry is used, where tangent points are set for 90 degrees at any valve lift or cam lift other than MID-LIFT, then these ratio changes are greater as they move through the arcing process. Since "ratio" is only measured at FULL LIFT, a rocker manufacturer can hide this excessive (wasteful) motion by setting the center-to-center dimensions so that their publicized ratio is close at full lift, then attach trick sounding terms to this wasteful and inefficient design, and bombard engine builders with lots of advertising that this is a virtue. It is not. Ratio, by definition, is only a snapshot of a specification which does not depict the real waste of cam acceleration and duration throughout the more critical stages of valve lift, even though net valve lift (and ratio) may be close to what is advertised.

To see the losses in valve acceleration from the extra crank degree rotation (area-under-the-curve), one needs to map the valve lift cycle in fine increments from closed to full open, then closed again, by measuring net valve lift directly at the valve and noting on graph paper the required degrees of crank rotation for each .010" or .020" (or whatever you choose) of valve lift to measure by. You can also do the test from the opposite perspective, and pick a specific crank degree step, like every 5 degrees (something small), and jot down the valve lift on graph paper. This latter example gives you a valve reaction perspective which will show lost valve lift at respective crank positions between one geometry test and another; while the former example gives you a cam reaction perspective, which illustrates the change of crank degrees for a given valve lift. You can do this test with two different pushrod lengths, back to back, of just .050" difference. Doesn't matter if you're too long or too short, just the results from the change should get your attention. You'll soon see for yourself that the closer you get to a true MID-LIFT geometry, the more quickly the valve will react to each degree of crank rotation.

Understand what is going on here: Rocker ratio is a mechanical division of the pushrod end into the roller tip (or shoe tip) end of the rocker. If you want to increase ratio, you shorten the pushrod dimension to the axis, or lengthen the roller tip dimension to the axis. But on these other rocker designs marketed under these misleading terms, there is no shifting, moving parts where the pushrod is at, nor does the roller tip move in and out within the rocker body to effect a different physical dimension to the axis (although some creative European engineers have patented some interesting concepts along these lines, but they don't apply to American OHV engines). So how can the rocker arm have variable ratios, or accelerating ratios, or quick lifting ratios? It simply has to do with the angle of the pivot point (shaft/trunnion axis) with the tangent points (axis) of the pushrod tip, and roller tip axis. By creating a 90 degree (or close to it) relationship in the CLOSED valve position, which is the angle when the greatest response velocity to cam motion will occur, the speed is "accelerated" in comparison to the other points of reference in the lift cycle, in fact, they will be decreasingly slower. In addition, the natural laws of physics will dictate that the arc from this starting point is increased proportionately, often doubling or tripling (or more) the in-and-out sweep.

Of course the other byproduct to this design is increased side loads on all contact points which endure this increased angle of attack; the valve stem in the valve guide, the pushrod tip into the adjusting screw or cup (often breaking them more easily), and let's not forget the additional side loads imposed on the bottom of the pushrod, where the resisting forces are now loading into the cam follower at a greater angle, and whip harmonics, which increases loads and wear, and drag coefficients within the tappet bore. None of this even touches the wasted cam information (mentioned above) through excessive motion in-and-out. Had enough? As noted elsewhere here, when you raise the rocker arm's pivot points to shift the geometry for promoting increased velocity off the seat, the rocker arm reverses the process as it closes. So if you've got a rocker design that sets geometry so that you've lifted the valve off the seat quickly, which inevitably slows down moving toward full lift, then when it has to close the valve will leave the full open position slowly and accelerate faster toward the closed position (i.e., the valve seat). This is the most destructive scenario you could wish for. You want to always set the valve down ON THE SEAT as gently as possible to minimize valve bounce. This is "Engine building 101." One term they've coined for this is called their "quick lift" technology; which is nothing more than "over-arcing" by another name. Don't be fooled! To read more details of these symptoms: See: "OVER-ARCING and UNDER-ARCING."

VELOCITY: (Generally) Means the rate of speed of any vehicle, item, component, fluid (including air) as measured in some constant value of time calculated by distance; or distance calculated by time. Angular Velocity is another derivative, and term sometimes used in rocker arm analysis with comparisons of rocker arm lengths and their affect with valve lifts. Port Velocity, Cam Velocity and Piston Velocity are all common terms used with engine builders.

VOLUMETRIC EFFICIENCY (VE): Is a percentage measurement of NET cylinder filling and operational throughput of airflow through the engine in comparison to its displacement. More simply stated, a cylinder having a volume of 100 cubic inches should theoretically be able to only hold 100 cubic inches. In so doing, it has achieved 100% "volumetric efficiency." In reality though, engines utilizing RAM TUNING for specific RPM ranges can optimize their efficiency beyond 100%, going over 110% or more. This is especially impressive when you keep in mind that displacement values of a cylinder are limited to the actual cam timing events of when the valves are opened and closed in relation to piston position. Since the valves are opening and closing before and after the piston's ultimate TOP and BOTTOM positions to anticipate delays in response of air flow momentum, inertia and response time of the flow depression traveling up the manifold runners that is created by the pistons, this superficial perspective of filling 100 cubic inches for 100% VE isn't as easy as implied, and the comparison isn't quite fair. See: NECP and COMPRESSION RATIO for additional info. NOTE: The tuning tricks mentioned here for optimizing VE applies to naturally aspirated engines, of peak high performance design utilizing maximum effort components with peak rpm's being reached. Supercharged, Turbo-charged and Nitrous-oxide induced systems realize over-charged VE through unnatural pressurizing of the cylinder. ^

WASTEGATE: (Turbocharger) The secondary passageway of a Turbocharger's exhaust driven hot section side, that provides either a fixed or adjustable relief of the engine's exhaust system pressure around the Turbocharger's Turbine, to limit unnecessary speed and Turbo pressure beyond a specifically designed amount for the engine in question.

WET FLOW: (Cylinder Heads) Refers to the relative air flow characteristics of a cylinder head's intake port in the more relevant "wet" state that an air/fuel mixture would have, in comparison to the "Dry Flow" characteristics that most flow bench study is done under, where only the dry, ambient air pulled through the port by the flow bench, and used to develop port shape and flow quality values. Although used for many years with specialized flow bench modifications or dedicated designs, it adds another dynamic that can not really be duplicated in a laboratory test stand condition, as it really acts in operation within a competitive engine. Violent inertia pulsations of the heavier liquid within the air flow jet stream from reversion waves; rapid temperature variations, port surface texture effect at these different speeds within the port, in most cases open up more questions to a comparison of dry flow analysis than what may get answered. Which is why most port development to this day is still constrained to dry flow testing, backed up by dyno testing, where more real world data acquisition can be taken and correlated to dry flow data.

WEAR PATTERN: (a.k.a. ROLL PATTERN) In rocker arms we use this term most often to describe and measure the actual wear pattern of the contact surface between the rocker arm's depressing roller or shoe tip upon the valve stem tip. Very specifically predictable under MID-LIFT geometry, it is actually a symptom of correct or incorrect geometry, NOT an accurate method for setting final geometry, because it's measuring the accuracy of a dimension that is at a right angle to its results. Unfortunately, some tools rely on this perspective to confirm correct geometry, lending to a false security.

WRONG GEOMETRY: One of the most confusing and most abused interpretations of setting rocker geometry relates to the "middle of the valve," which I have noted for many years has NOTHING to do with setting correct rocker geometry. Yet here we are, 2009, and after more than three years (2006) since I made the connection of this term to their web site link, we still have one of the most well known, most advertised "brand names" still throws high octane fuel on the fire of confusion. Worse, they use our Registered Trademark term, MID-LIFT® to confuse people even more (even though they left out the hyphen - lawyer's trick). I don't often refer people to other company's web sites, but we sometimes make an exception. The technical illustration from our old friends in Memphis is a great example of exactly what you DON'T want to do. Follow it, and you will have the ultimate example of total over-arcing. If we wanted to draw an example that was an antonym of the REAL MID-LIFT, this is what we would do! Wrong geometry is ironically very fitting as the LAST TERM of our Terms & Definitions page. After you've jumped over to check this out, don't forget to use the <---BACK button on your browser to return; you don't want to get lost back in the logic of the 1960s!  ^