Rod Length Relationships
You are invited to participate in this attempt to understand a part of internal combustion engines. I invite any/all criticisms, suggestions, thoughts, analogies, etc.-- written preferred, phone calls accepted from those too feeble or who have arthritis. Contributors are invited to request special computer printouts for specific combinations of interest to them.
In general, most observations relate to engines used for some type of competition event and will in general produce peak power higher than 6000 RPM with good compression ring seal as defined by no more than 3/16 CFM blowby per cylinder.
Short Rod is slower at BDC range and faster at TDC range.
Long Rod is faster at BDC range and slower at TDC range.
I. LONG ROD
A. Intake Stroke -- will draw harder on cyl head from 90-o ATDC to BDC.
B. Compression Stroke -- Piston travels from BDC to 90-o BTDC faster than short rod. Goes slower from 90-o BTDC to TDC--may change ign timing requirement versus short rod as piston spends more time at top. However; if flame travel were too fast, detonation could occur. Is it possible the long rod could have more cyl pressure at ie. 30-o ATDC but less crankpin force at 70-o ATDC. Does a long rod produce more efficient combustion at high RPM--measure CO, CO2? Find out!!
C. Power Stroke -- Piston is further down in bore for any given rod/crank pin angle and thus, at any crank angle from 20 to 75 ATDC less force is exerted on the crank pin than a shorter rod. However, the piston will be higher in the bore for any given crank angle from 90-o BTDC to 90-o ATDC and thus cylinder pressure could be higher. Long rod will spend less time from 90-o ATDC to BDC--allows less time for exhaust to escape on power stroke and will force more exhaust out from BDC to 90-o BTDC. Could have more pumping loss! Could be if exhaust port is poor, a long rod will help peak power.
D. Exhaust Stroke -- see above.
II. Short Rod
A. Intake Stroke -- Short rod spends less time near TDC and will suck harder on the cyl head from 10-o ATDC to 90-o ATDC the early part of the stroke, but will not suck as hard from 90-o to BDC as a long rod. Will require a better cyl head than long rod to produce same peak HP. Short rod may work better for a IR or Tuned runner system that would probably have more inertia cyl filling than a short runner system as piston passes BDC. Will require stronger wrist pins, piston pin bosses, and connecting rods than a long rod.
B. Compression Stroke -- Piston moves slower from BDC to 90-o BTDC; faster from 90-o BTDC to TDC than long rod. Thus, with same ign timing short rod will create less cyl compression for any given crank angle from 90-o BTDC to 90-o ATDC except at TDC. As piston comes down, it will have moved further; thus, from a "time" standpoint, the short rod may be less prone to detonation and may permit higher comp ratios. Short rod spends more time at the bottom which may reduce intake charge being pumped back out intake tract as valve closes--ie. may permit longer intake lobe and/or later intake closing than a long rod.
C. Power Stroke -- Short rod exerts more force to the crank pin at any crank angle that counts ie.--20-o ATDC to 70-o ATDC. Also side loads cyl walls more than long rod. Will probably be more critical of piston design and cyl wall rigidity.
D. Exhaust Stroke -- Stroke starts anywhere from 80-o to 110-o BBDC in race engines due to exhaust valve opening. Permits earlier exhaust opening due to cyl pressure/force being delivered to crank pin sooner with short rod. Requires a better exhaust port as it will not pump like a long rod. Short rod has less pumping loss ABDC up to 90-o BTDC and has more pumping loss from 90-o BTDC as it approaches TDC, and may cause more reversion.
A. Rod Length Changes -- Appears a length change of 2-1/2% is necessary to perceive a change was made. For R & D purposes it appears a 5% change should be made. Perhaps any change should be 2 to 3%--ie. Ignition timing, header tube area, pipe length, cam shaft valve event area, cyl head flow change, etc.
B. Short Rod in Power Stroke -- Piston is higher in the bore when Rod-Crank angle is at 90-o even though at any given crank angle the piston is further down. Thus, at any given "time" on the power stroke between a rod to crank pin angle of 10o and ie. 90-o, the short rod will generate a greater force on the crank pin which will be in the 70-o to 75-o ATDC range for most engines we are concerned with.
C. Stroke -- Trend of OEM engine mfgs to go to longer stroke and/or less over square (bore numerically higher than stroke) may be a function of L/R. Being that at slower engine speeds the effect of a short rod on Intake causes few problems. Compression/Power Stroke should produce different emissions than a long rod. Short rod Exhaust Stroke may create more reversion--EGR on a street engine.
D. More exhaust lobe or a earlier exhaust opening may defeat a longer rod. I am saying that a shorter rod allows a earlier exhaust opening. A better exhaust port allows a earlier exhaust opening.
E. Definition of poor exhaust port. Becomes turbulent at lower velocity than a better port. Flow curve will flatten out at a lower lift than a good port. A good exhaust port will tolerate more exhaust lobe and the engine will like it. Presuming the engine has adequate throttle area (so as not to cause more than 1" Hg depression below inlet throttle at peak power); then the better the exhaust port is, the greater the differential between optimum intake lobe duration and exhaust lobe duration will be--ie. exh 10-o or more longer than intake Carbon buildup will be minimal if cyl is dry.
Short Rod -- Min Rod/Stroke Ratio -- 1.60 Max Rod/Stroke Ratio -- 1.80
Long Rod -- Min Rod/Stroke Ratio -- 1.81 Max Rod/Stroke Ratio -- 2.00
Any ratio's exceeding these boundaries are at this moment labeled "design screw-ups" and not worth considering until valid data supports it.
Contributors to Date: Bill Clemmons, Jere Stahl
Connecting Rod Length Influence on Power
by William B. Clemmens
A spark ignition (SI) engine and a steam engine are very similar in principle. Both rely on pressure above the piston to produce rotary power. Pressure above the piston times the area of the bore acts to create a force that acts through the connecting rod to rotate the crankshaft. If the crankshaft is looked at as a simple lever with which to gain mechanical advantage, the greatest advantage would occur when the force was applied at right angles to the crankshaft. If this analogy is carried to the connecting rod crankshaft interface, it would suggest that the most efficient mechanical use of the cylinder pressure would occur when the crank and the connecting rod are at right angles. Changing the connecting rod length relative to the stroke changes the time in crank angle degrees necessary to reach the right angle condition.
A short connecting rod achieves this right angle condition sooner than a long rod. Therefore from a "time" perspective, a short rod would always be the choice for maximum torque. The shorter rod achieves the right angle position sooner and it does so with the piston slightly farther up in the bore. This means that the cyl pressure (or force on the piston) in the cylinder is slightly higher in the short rod engine compared to the long rod engine (relative to time).
ROD LENGTH RELATIONSHIPS*
(with Crank @ 90 deg ATDC)
Piston Position Crankpin/Rod Angle
|Stroke||Rod Length||Rod Angle||from TDC||ATDC|
ROD LENGTH RELATIONSHIPS with CRANKPIN/ROD centerline @ 90o @ 7500 rpm
|Stroke||Rod Length||Rod Angle||Piston Distance||Crank Angle||Piston Accel|
*data from Jere Stahl
Another concern in selecting the rod length is the effects of mechanical stress imposed by increasing engine speed. Typically, the concept of mean piston speed is used to express the level of mechanical stress. However, the word "mean" refers to the average speed of the piston in going from the top of the bore to the bottom of the bore and back to the top of the bore. This distance is a linear distance and is a function of the engine stroke and engine speed, not rod length. Therefore, the mean piston speed would be the same for each rod length listed in Table 1.
Empirical experience; however, indicates that the mechanical stress is less with the longer rod length. There are two reasons for these results. Probably the primary reason for these results is that the profile of the instantaneous velocity of the piston changes with rod length. The longer rod allows the piston to come to a stop at the top of the bore and accelerate away much more slowly than a short rod engine. This slower motion translates into a lower instantaneous velocity and hence lower stresses on the piston. Another strong effect on mechanical stress levels is the angle of the connecting rod with the bore centerline during the engine cycle. The smaller the centerline angle, the less the side loading on the cylinder wall. The longer rod will have less centerline angle for the same crank angle than the shorter rod and therefore has lower side loadings.
Classical textbooks by Obert ( ) and C.F. Taylor ( ) provide little guidance on the rod length selection for passenger or commercial vehicles other than to list the ratios of rod length to crank radiuses that have been used by various engine designs. Race engine builders using production blocks have done quite a bit of experimentation and have found many drivers are capable of telling the difference and making clear choices along with similar results from motorcycle flat track racers/builders.
Because of recent developments in computer modeling of the engine cycle by R.D. Rabbitt ( ), another factor may be critical in selecting a given connecting rod length. This new factor is the cylinder head flow capability versus connecting rod length over stroke ratio (l/r) versus engine speed. To understand this relationship, let us first review previous techniques used to model air flow during the engine cycle which as Rabbitt points out is founded on principles initiated in 1862 and refined in 1920. These theories are documented in Taylor's textbook ( ). To calculate air flow throughout the cycle these models use such parameters as mean or average inlet mach number for the port velocity and an average inlet valve discharge coefficient which compensate for valve lift and duration. In these models a control volume is used to define the boundaries of the combustion chamber. The air flow determined by the previous parameters crosses this boundary to provide air (and fuel) for the combustion process within the control volume.
However, this control volume has historically been drawn in a manner that defines the boundaries of the combustion chamber in the area of the inlet and exhaust valves as if the valves were removed from the cylinder head (ie. a straight line across the port). With the valves effectively removed, the previously mentioned average port flow and valve discharge coefficient (ie. valve restriction) are multiplied within current computer models to quantify the air flow (and fuel) delivered for each intake stroke. But, as Rabbitt points out, this approach totally ignores the effect of the air flow direction and the real effect of valve lift on the total air flow that can be ingested on each intake stroke.
Rabbitt reaches two important conclusions from his study. One, because of the direction of the air flow (angle and swirl) entering the combustion chamber, three dimentional vorticies are set up during the intake stroke. Two, that above a certain piston speed, density of the mixture at the piston face is a function of valve geometry and valve speed. Rabbitt further discusses the effect of the first conclusion as it relates to the mass of air that is allowed to flow through the port and by the valve. Vorticies can exhibit different characteristics and in general conform to two general types--large scale bulk vorticies that could be described as smooth in nature and small scale eddies that are highly turbulant.
If one can consider that the vacuum produced by the piston on its downward travel to be the energy that causes the air to flow through the port when energy losses throughout the intake tract (including losses at the valve) are at a minimum, the flow delivered to the chamber will be maximized. If the area between the piston face and the valve is also included in the consideration of flow losses, the effect of the type of vorticies created can be more easily understood. Large scale bulk vorticies comsume less energy than highly turbulent eddy vorticies. Thus, more of the initial energy from the piston's downward movement is available at the port-valve-combustion chamber interface with which to draw the intake charge into the chamber. Small scale eddies eat up energy which reduces the amount of the initial energy that reaches the port-valve-combustion chamber interface which in turn, reduces the port flow.
Rabbitt's second conclusion follows that at some higher piston speed, the vorticies within the combustion chamber (which are assumed to be large scale bulk type at low speeds) transition from the bulk type to the small scale eddy type. At this point the flow into the combustion chamber ceases to increase in proportion to increases in engine speed. It is theorized that this flow transition point can be observed on the engine power curve as the point at which the power curve begins to fall off with increasing engine speed.
As indicated earlier, piston speed is normally viewed as mean or average piston speed. Thus for a given engine, the mean piston speed increases as the rotational engine speed increases. However, in Rabbitt's model the piston speed of concern is the instantaneous piston speed during the intake stroke near TDC. For any given engine, changing the rod length to stroke (l/r) ratio changes the instantaneous piston speed near TDC. For the purposes of flow visualization, the type of vortex formed should not care whether a given instantaneous piston speed had been achieved by a given rotational speed or changing the (l/r) ratio and operating at a new rotational speed. As long as the instantaneous piston velocity is the same, the type of vorticies formed should be the same and the amount of air inducted into the cylinder should be the same.
If other factors influenced by rotational speed such as the time distance between slug of intake air flow and valve opening rates relative to the acceleration of the air slugs were ignored, one should be able to predict the location (RPM) of the peak power as a result of a change in the (l/r) ratio. Note, that even though power is a funtion of air flow and air flow should be roughly constant for the same instantaneous piston speed (neglecting the afore mentioned factors), the power may not be the same because of the lever arm effect between the crank radius and the connecting rod. (As we noted earlier, the shorter rod should have the advantage in the lever arm effect.)
In reality, the analysis must be viewed by stroke (ie intake, compression, exhaust, power) the selection of exhaust valve opening time combined with the exhaust system backpressure and degree of turbulance the exhaust port experiences. If the exhaust port has good turbulance control then you may run a shorter rod which allows you to use more exhaust lobe which reduces pumping losses on the exhaust stroke.
My engine theories are based on a heavy research effort since the late 80's. I have studied engineering texts, selected magazine and internet articles, as well as talking with engine builders, racers, and cam specialists. The people I have had the priviledge to exchange ideas with, as well as learn a few really cool things from, have proven to be invaluable. I am a fan of most all engine combinations that are built within the structured hardware limitations of class racing. As a mechanical engineer, I do not subscribe to the idea that a motor works because 'it just does' or from mystic voodoo black magic than only a few understand. An engine is a mathematically quantifiable system and should be treated and/or modified in this light. However, the problem that will always exist with math calculations is: garbage in equals garbage out. An accurate model is required to give accurate results.
I am an engine enthusiast to say the least. Cubic dollars have prevented me from furthering my involvement in race engines other than my own bracket car and engines in some of my friends’ cars. My ideas come from my studies as a mechanical engineer and reading technical data of any shape or form pertaining to engine theory.
In 1988, I began formulating a simple calculation that would enable me to approximate the power level of a particular engine size and family, based on intake flow and rpm. At that time, I was using maximum cylinder head flow in my calculations and assigning compression ratios to a particular cam duration. Since then, I have quantified the role of port volume/area as well as cylinder pressure in the engine parameters. My engine calculations are continually evolving and are getting fairly detailed, however most of the last few years (since '95) have been spent formulating valve timing calculations. After several years of relying on curve-fit equations for rpm versus duration, I came to the realization that this was not reality. The equations provided a very convenient baseline; but at some point, valve timing must become a collection of opening and closing points, not an imposed duration value.
I am including general ideas on cylinder heads, cams, and to a lesser degree (but no less important) intake manifolds and headers because these components are among the most common bolt-on parts, and therefore the most commonly mismatched parts.
Cylinder Head Requirements:
Maximum runner flow is somewhat meaningless without the total flow curve throughout the complete lift range of the cam. Flow occurs from the time the valve is opened and until it is closed. If the flow values do not match the needs of the engine throughout the lift range, then the ability to make good power is gone everywhere but at that those specific matching valve lifts. The end result is poor cylinder filling because of flow inadequacies. If you apply this power loss to every cylinder over the whole rpm range, it becomes quickly apparent that the power output has taken a significant hit. In short, the cylinder is not adequately being filled and the cylinder pressure is suffering. Inadequate cylinder pressure adds up to reduced torque.
A useful relationship to keep in mind is the runner's flow per cross-sectional area. If the flow area is calculated over the entire lift range and divided by port's cross-sectional area, the resulting proportion is a great benchmark for comparing runner effectiveness. Using a method like this helps to average the total flow capability, over the complete valve lift range, and yields a more realistic look at the power output based on intake flow capability.
The intake manifold’s ability to affect the cylinder head’s flow capability cannot be overlooked. It is beneficial to have flow data with the intake manifold passage attached to the head, as well as the intake manifold runner volumes and cross-section areas. The combination of the two port runners make up the engine’s intake port, not just the 3-6 inch length cast into the cylinder head.
This is an area where my theories may deviate from what I classify as normal thinking. Correct flow quality must be maintained within the cam’s lift range, just as on the intake port. I think of the exhaust port as an extension of the valve seat/bowl geometry and the header tube as the actual exhaust runner. Because of this relationship, the header tube should be installed when flow data is taken on the exhaust, or at least a tube stub simulating the header diameter and transition with enough length to introduce flow losses like the actual header. All modifications to the port cross-section should be in the interest of maintaining an effective transition into the header tube, as well as increasing flow. The exhaust runner will behave differently than the intake runner because it is a tube with constant cross-section (usually). Short of using inertial wave effects to improve the overall port effectiveness, optimizing the seat geometry, valve bowl area, and effectively transitioning the flow to the header tube is about all than can be done because the maximum flow capability is dictated by the port's cross-sectional area. The header primary tube is the critical dimension in the exhaust path and therefore controls maximum flow capability. This total flow is even effected by the exhaust system as a whole.
Runner Volume and Cross-Sectional Area:
Runner Volume, like maximum runner flow, is meaningless without the complete picture. The port's cross-sectional area is the real parameter to characterize. Compare the small block Chevrolet intake port volume to the Ford's. A 200cc runner in the Chevrolet head can be considerably different than the same volume in a Ford small block head. The stock Chevrolet runner is about and inch longer, which for a given port volume will yield a smaller port cross section. If runner volume is the only quantity known, then the port length has to be measured to calculate an average cross-sectional area or the port must be disected to find the most representative cross-sectional area. An engine’s torque peak is directly related to the intake and exhaust runner’s cross-sectional area. Large cross-sections, in proportion to the cylinder volume will produce higher rpm torque peaks.
Along with port cross-sectional area, is the requirement of the port to produce flow velocities indicative of an efficient port design. Therefore a minimum flow rate exists for a particular cross-sectional area to maintain this velocity. It is important to realize that forced induction and high volumetric efficiency motors have breathing capabilities beyond their cubic inch limitations and will follow these trends a little differently. In these special cases, the engine components must reflect this increased breathing capability. In short, if an engine has substantial intake flow capability, then the exhaust side must be equally sized, regardless of the engine's cubic inch displacement or horsepower rating.
The camshaft is the brain of the motor. It controls everything that happens. Because of this, I have basically abandoned all cam catalog applications and go directly to the lobe specification sections. An off-the-shelf cam is a compromise, at best, to get optimum valve timing into a wide range of engine set-ups. Because of this marketing scheme, you see information such as minimum compression ratio, generic intake and exhaust requirements, and rpm ranges for particular families or series of grinds.
Look at two engines, each with the same displacement and geometry, one of which has very large cylinder head runners with very high flow capability and the other with virtually stock heads. Each engine is outfitted with identical intake manifolds and headers/exhaust. The two engines will exhibit very different requirements for valve timing for similar performance and rpm capability. This is the sort of information that is not addressed in most catalog offerings. Equally different, are two engines outfitted identically, but with differing cubic inch displacements. In these examples, assigning camshaft specifications based on broad generalized ideas used by most off-the-shelf marketing techniques can fall short of realistic engine requirements. A camshaft selected on rpm alone is not going to match the engine's performance needs unless the camshaft was designed around a specific set of engine parameters, and your set-up matches it identically.
The cam should always be the last thing chosen on a motor set-up. The ramp profiles can crutch low-lift flow issues, intake and/or carburetor restrictions, and exhaust inadequacies. Overlap quantities are selected to increase cylinder scavenging, aid in low or high flow velocity situations, or to meet special idle or driveability requirements. Intake closing points tailor rpm requirements. Exhaust opening points will reflect the exhaust system capability. All of these characteristics paired up with the vehicle's gear ratios will define the overall package's effectiveness.
: Specific durations and lobe separations have not been mentioned. Engines do not read cam catalogs or magazine articles. An engine makes power from cylinder pressure, port velocity, and rpm. Valve openings and closings are the real numbers to be concerned with. We always speak in terms of duration and lobe separation; but keep in mind, those values are byproducts of the valve openings and closings.
One of my early theories was to apply a minimum static compression ratio to a particular camshaft duration. This was only a part of the puzzle. Compression ratio helps, but what about the low compression motors that make big time horsepower? Cylinder pressure and port velocity are the key players in determining power output. The camshaft is the controlling factor for both of these parameters. By optimizing the timing events within the crank cycle for the required rpm band, the cylinder pressure can be tailored to compliment the engine geometry and components. Typically, the goal is to maximize (or atleast optimize) the cylinder pressure within the engine's intended rpm range.
The media globally labels the Lobe Separation Angle (LSA) as a major selection criteria for choosing a camshaft. Narrower LSA cams are usually touted as having poor idle and driveability characteristics. Part of the reason they have such limited usage guidelines is because of their supposed 'peaky' powerband characteristics. If the cam design is allowed to compliment high port velocity and cylinder pressure for the intended rpm range, then idle and driveability will be no worse than a wider LSA cam used in its correct application, for the same rpm range. In reality, the LSA is a relative value and does not pertain to the intended usage of the engine combination. Poor idle and driveability characteristics come from mismatched components and from the purchase of components with design requirements way out of line from the engine’s needs.
No text on cam design can be complete without discussing overlap. Overlap has been a 'buzzword' around camshaft discussions since hotrodding began. Technically, overlap is the point in the crank cycle at which there is charge exchange during the intake opening and exhaust closing at TDC. It can be an extremely misunderstood process. Like any engine subject: more is better, less is better, it depends on whoever's musings you read or listen to. Which is true, why is there so much contradiction? The media has applied guidelines, like the lobe separation theme, that are unconfounded. The ideas may not be wrong, it's just that the complete story is not told.
Overlap is required to aid in the air exchange, increase the overall breathing capacity of an engine, and tailor rpm band width. Two terms generally associated with overlap are 'scavenging' and 'reversion'. The situation exists that if there is too much pressure in the exhaust, due to flow restriction or inertia wave pulses, then it is possible to reverse the flow in the intake port (reversion). Likewise if the exhaust pressure is too low, creating a vacuum, it is possible for the incoming charge to be sucked past the chamber and go out the exhaust (over-scavenging). The goal is for the pressure differential and charge momentum to be just enough so that the intake charge is helped into the chamber, but the exhaust valve closes just in time to prevent excessive intake charge from being carried out with the exhaust. The key players in determining an engine's overlap requirements are the port cross-sectional areas, flow capability, rod/stroke ratio, static compression, and rpm.
Overlap is a quantity of flow per degree of crank rotation; not so much an angular value based on the engine's intended operating environment. When you look at overlap as a quantity of air exchange per crank rotation or piston displacement, it becomes apparent why very high flow capability motors use less overlap and therefore wider LSA's. The angular measurement is less, but the actual air exchange is still occurring.
I did not get into any of the Inertial Wave Tuning effects theory. I am not sure I have an engineering grasp on the topic or the assumptions that can be made to account for the temperature and pressure variations. The topics I have spoken about are in the interest of maximizing power output per a given engine set-up. These topics along with the inertial wave tuning are vitally important and play a major role in the engine’s power making ability. But, I believe that by applying the relationships of the port size, flow, and rpm, you can at least be in the game even if you are not hitting home runs. Based on these ideas, it appears that there may be power limitations for an engine set-up by omitting inertial effects. I believe this is true. However, a good baseline set-up is going to provide a much better starting point. Designing an engine set-up around inertial effects without careful consideration of the flow parameters (cross-sections, flow, velocity, valve events, cylinder volumes, etc) is pointless.
I think the ‘Class’ racers/builders have the best knowledge of engine set-ups, as opposed to the builders that have the ‘eternal quest’ for maximum power in an unlimited environment of hardware and components. I am referring to the ‘class’ engine builders in drag racing, stock car, and sports cars that are forced to work within a particular hardware ‘envelope’. Once you get knowledge of how to make a set-up work within an ‘envelope’, then all buildups will benefit because of the understanding of what various components do and what the limitations are.
Overlap and Compression- A very common idea, although for the most part incorrect, is that overlap bleeds off compression. Overlap, by itself, does not bleed off compression. Overlap is the angle between the exhaust closing and intake opening and is used to tune the exhaust's ability draw in additional intake charge as well as tuning idle vacuum and controlling power band width. Cylinder pressure is generated during the compression cycle, after the intake valve has closed and before the exhaust opens. Within practical limits, an early intake closing and late exhaust opening will maintain the highest cylinder pressure. By narrowing the Lobe Seperation Angle 'LSA' for a given lobe duration, the overlap increases, but the cylinder pressure can be increased as well. Thus cylinder pressure/compression can actually increase in this scenario, by the earlier intake closing and later exhaust opening. By increasing duration for a given LSA, the overlap will increase, the intake closing will be delayed, and the exhaust opening will occur earlier. This will decrease cylinder pressure, but the decrease/bleed-off of compression is not due to the overlap, it is due to the intake closing and exhaust opening events.
Adjusting Lash on Mechanical/Solid Cams- If valve lash changes significantly over time, then something is wrong. Cam wear is very slight, along the order of .002 or less. If the lash setting changes more than .005 then there has been a component failure (loosened hardware or actual mechanical failure). Lash settings should be taken/adjusted at the same temperature and same order as the previous or original setting. This is the only way to rule out expansion/contraction of the components from temperature changes. This temperature delta is usually the culprit of most valve lash dilemmas. At initial start-up and break-in of a new set-up: cam, lifters, rockers, pushrods, valve job, etc., the lash may move around during the break-in procedure and for a short time after. This is because all the parts are seating into their new wear patterns. Once this occurs, the lash setting should stay steady.
Hydraulic Lifter PreLoad- Hydraulic lifters are intended to make up for valvetrain dimensional differences as well as providing a self-adjusting method of maintaining valve lash, or rather the lack of. By setting the valvetrain so the lifter plunger is depressed slightly, the lifter is able to compensate for these differences, making a convenient hassle-free valvetrain set-up. For performance applications, lifter preload is not needed or wanted. As rpm's increase, the lifter has a tendency to bounce over the back of the lobe as it comes back down from the maximum lift point. The pressurized oil fills the lifter body to account for this bouncing. Eventually, after several engine revolutions, the oil can completely fill the lifter body and the plunger will be pushed up to its full travel (pump-up). Higher oil pressures can amplify this problem. With the lifter pre-loaded, this can cause a valve to run off it's seat and can cause piston clearance issues if and when pump-up occurs. By setting the valvetrain at 'zero' preload, lifter pump up is eliminated and in most cases, the cam will rev higher. Ford tech articles in late 60's actually urged 'stock' class racers to run .001-.003 lash on hydraulic cams.
Piston To Valve Clearance- Piston clearance is a function of lobe geometry and phasing to the piston. Cam lift should not be a deciding a factor in clearance issues. Valves will hit the piston in the overlap period, while exhaust is closing and intake is opening. Exhaust clearance problems will typically occur just before TDC and intake just after TDC, not at max lift. Some cylinder head venders and other component manufacturers advertise a max duration or lift before clearance issues arise. This is very misleading. Maximum safe duration is a totally bogus value, and is completely worthless without knowing anything about the ramp rates or actual timing/phasing events of the installation. At least with maximum safe lift, the vendor can a apply a rediculously fast ramp at a very early opening/closing and arrive at a somewhat meaningful measurement, but without knowing the design specifics the information is still next to useless.
Custom Ground Camshafts- When the performance of a particular engine combination is wanted to be optimized, the camshaft design parameters are calculated from the engine and vehicle specifications to perform within specific conditions. Let me emphasize that last statement, 'within specific conditions!'. In no way was total maximum power for the engine implied. The intent is to maximize performance within the intended design parameters. If that means taking a pro-stock motor and wanting to run it from 2000-5000 rpm, then so be it.
The camshaft's seat timing events, ramp rate, and lift are directly related to the intake and exhaust flow capabilities, crankshaft geometry, static compression, rpm range, as well as other criteria. A camshaft selected in this manner, becomes personalized to that particular engine combination. Usually a custom grind is selected as an intake lobe and exhaust lobe with a particular phasing to each other (lobe separation angle, LSA) and sometimes a specified amount of advance or retard is built in. Although, it could easily end up having completely reengineered lobe characteristics, requiring new lobe masters with specialized ramp requirements. It is possible for an off-the-shelf camshaft to be a classified as a 'custom'. If the cam design is calculated for a particular combination and an off-the-shelf part number fits the bill, then for all practical purposes that part number is a 'custom' cam (but only for that particular set-up).
Typically, cam catalogs do not specifically list custom ground camshafts, because the possibilities are endless. They stick to particular series or families of camshafts. The superstock grinds come closest to an off-the-shelf grind that is truly optimized for a combination. There will be small differences due to header sizes and engine builder's 'secrets, but usually the catalogs are pretty close to a good baseline. Likewise, brand to brand, the grinds will be very similar because of the 'class' dictated combinations and the flow characteristics being so well documented
Degreeing Camshafts- There is no special magic involved for degreeing a camshaft during installation, but this is not the same thing as random advancing, retarding, or installing the gears 'lined up'. Degreeing a camshaft involves definite known values for valve events. Typically this is specified as an Intake Centerline or as opening/closing events at specific lobe lifts. This is done to insure the cam is installed per specific requirements, such as a recommendation from an engine builder or the vendor's data sheet for that camshaft grind. Manufacturing tolerances and shop practices do not guarantee that the cam matches the data sheet, when installed at crank gear 'zero'. The cam will usually need to be advanced or retarded to the correct location. If it is correct, at crank gear 'zero', then the cam has still been degreed. It just did not require any additional tweaking to meet the requirements. This is what degreeing a cam is all about; the verification of the installation. A common mis-used term is the 'straight-up' installation. Typically this is described as installing the cam at crank gear 'zero'. This is 100% wrong. Straight-up refers to the Intake and Exhaust Centerlines being the same. In other words the cam will have no advance or retard at the installation, regardless of the amount of advance/retard ground in by the vendor. In reality, the cam may have to be advanced or retarded (from crank gear 'zero') significantly to arrive at a straight-up installation.
Exhaust System Diameter and Engine Horsepower- A popular idea is to select/size the exhaust system components to the engine's horsepower output. This idea typically attributes a header diameter or an exhaust system diameter to a particular horsepower level. To resolve this, look at how an engine operates and consider one cylinder. The cylinder will move a volume of air based on its crankshaft geometry, rpm, and sealing capability. The amount of air that can enter the cylinder is dependant on the intake flow capability, crank geometry, rpm, and valve timing as a minimum consideration. Likewise, the amount of air that exits the cylinder is dependent on the same characteristics.
An engine's output is usually thought of in terms of horsepower. Actually, an engine produces torque, and the horsepower is calculated through a units conversion. The amount of torque an engine can produce is directly related to the amount of cylinder pressure generated. This is all affected by the same previous characteristics (intake and exhaust capability, crank geometry, rpm, valvetiming, etc). So basically an engine's power output is about air exchange capability. Using this line of thinking, look at the exhaust path again. The exhaust system is more reflective of the engine's ability to move air, as opposed to horsepower numbers. Engine output does not address the breathing aspects of the engine and is probably not a good rule to use for exhaust sizing.
There is a very good reason that tuners/engineers/specialist have attempted to assign exhaust to intake relationships around 70-80% for a typical natural aspirated set-up. In non-detailed terms, it is a range that offers a good balance for power capability. Other relationships, such as 1:1, are used and they work very well, but these methods have to be applied and tuned for very specific circumstances. This relationship does not stop on the flow bench, it goes all the way from the intake path opening to the exhaust system termination. In short, try to maintain exhaust sizes that are inline with the intake capability. Also, do not stop your analysis at the intake and exhaust paths. If the engine already has the camshaft, look at the valve events. If the specs favor a restricted exhaust (indicated by early and wider exhaust openings with wider lobe separation angles), then size it accordingly by using exhaust components with smaller cross-sections. If the valve timing specs favor the intake, then the engine needs some serious exhaust flow capability which is only possible with larger cross-sections.
This section was written with natural aspirated combinations in mind. However, by using the 'air exchange' rationale, it becomes apparent why forced induction engines typically benefit from increased exhaust flow capability. Also, look at the nitrous combinations. The intake system remains virtually unchanged, yet with the major increases in cylinder pressure it acts like a substantially larger engine on the exhaust side, requiring earlier exhaust openings and/or higher exhaust flow capability.
Pushrod Length- Incorrect pushrod length can be detrimental to valve guide wear. Most sources say that centering the rocker contact patch on the valve stem centerline at mid valve lift is the correct method for determining the optimum pushrod length. This method is wrong and can actually cause more harm than good. The method only applies when the valvetrain geometry is correct. This means that the rocker arm lengths and stud placement and valve tip heights are all perfect. This is rarely the case. To illustrate this, think of the valve angle and the rocker stud angle. They are usually not the same. If a longer or shorter valve is installed, then the relationship of the valve tip to the rocker stud centerline has changed. Heads that have had multiple valve jobs can also see this relationship change. Note, the rocker length (pivot to tip) remains unchanged, so the rocker contact patch will have to move off the valve centerline some particular distance for optimum geometry to be maintained.
The optimum length, for component longevity, is the length that will give the least rocker arm contact area on the valve stem. In other words the narrowest wear pattern. This assures that the relationship is optimized and the rocker is positioned at the correct angle. This means that the optimum rocker tip contact point does not necessarily coincide with the valve stem centerline, and probably will not. What is the acceptable limit for being offset from the valve stem centerline? That will depend on the set-up. A safe margin to strive for is about +/-.080" of the centerline of an 11/32 diameter valve stem. This means that no part of the wear pattern should be outside of this .160" wide envelope. As the pushrod length is changed, the pattern will change noticeably. As the geometry becomes closer to optimum, the pattern will get narrowest. If the narrowest pattern is too far from the valvestem centerline, then the valve to rocker relationship has to be changed. In this case, valve stem length will need to change.
Rod Lengths/Ratios: Much ado about almost nothing. From Isky Cams
Why do people change connecting rod lengths or alter their rod length to stroke ratios? I know why, they think they are changing them. They expect to gain (usually based upon the hype of some magazine article or the sales pitch of someone in the parts business) Torque or Horsepower here or there in rather significant "chunks". Well, they will experience some gains and losses here or there in torque and or H.P., but unfortunately these "chunks" everyone talks about are more like "chips".
To hear the hype about running a longer Rod and making more Torque @ low to mid RPM or mid to high RPM (yes, it is, believe it or not actually pitched both ways) you'd think that there must be a tremendous potential for gain, otherwise, why would anyone even bother? Good question. Let's begin with the basics. The manufacture's (Chevy, Ford, Chrysler etc.) employ automotive engineers and designers to do their best (especially today) in creating engine packages that are both powerful and efficient. They of course, must also consider longevity, for what good would come form designing an engine with say 5% more power at a price of one half the life factor? Obviously none. You usually don't get something for nothing - everything usually has its price. For example: I can design a cam with tremendous high RPM/H.P. potential, but it would be silly of me (not to mention the height of arrogance) to criticize the engineer who designed the stock camshaft. For this engine when I know how poorly this cam would perform at the lower operating RPM range in which this engineer was concerned with as his design objective!
Yet, I read of and hear about people who do this all the time with Rod lengths. They actually speak of the automotive engine designer responsible for running "such a short Rod" as a "stupid SOB." Well, folks I am here to tell you that those who spew such garbage should be ashamed of themselves - and not just because the original designer had different design criteria and objectives. I may shock some of you, but in your wildest dreams you are never going to achieve the level of power increase by changing your connecting rod lengths that you would, say in increasing compression ratio, cam duration or cylinder head flow capacity. To illustrate my point, take a look at the chart below. I have illustrated the crank angles and relative piston positions of today's most popular racing engine, the 3.48" stroke small block 350 V8 Chevy in standard 5.7", 6.00", 6.125" and 6.250" long rod lengths in 5 degree increments. Notice the infinitesimal (look it up in the dictionary) change in piston position for a given crank angle with the 4 different length rods. Not much here folks, but "oh, there must be a big difference in piston velocity, right?" Wrong! Again it's a marginal difference (check the source yourself - its performance calculator).
To hear all this hype about rod lengths I'm sure you were prepared for a nice 30, 40, or 50 HP increase, weren't you? Well its more like a 5-7 HP increase at best, and guess what? It comes at a price. The longer the rod, the closer your wrist pin boss will be to your ring lands. In extreme situations, 6.125" & 6.250" lengths for example, both ring and piston life are affected. The rings get a double whammy affect. First, with the pin boss crowding the rings, the normally designed space between the lands must be reduced to accommodate the higher wrist pin boss. Second, the rings wobble more and lose the seal of their fine edge as the piston rocks. A longer Rod influences the piston to dwell a bit longer at TDC than a shorter rod would and conversely, to dwell somewhat less at BDC. This is another area where people often get the information backwards.
In fact, this may surprise you, but I know of a gentleman who runs a 5.5" Rod in a 350 Small Block Chevy who makes more horsepower (we're talking top end here) than he would with a longer rod. Why? Because with a longer dwell time at BDC the short rod will actually allow you a slightly later intake closing point (about 1 or 2 degrees) in terms of crank angle, with the same piston rise in the cylinder. So in terms of the engines sensitivity to "reversion" with the shorter rod lengths you can run about 2-4 degrees more duration (1-2 degrees on both the opening & closing sides) without suffering this adverse affect! So much for the belief that longer rod's always enhance top end power!
Now to the subject of rod to stroke ratios. People are always looking for the "magic number" here - as if like Pythagoras they could possibly discover a mathematical relationship which would secure them a place in history. Rod to stroke ratios are for the most part the naturally occurring result of other engine design criteria. In other-words, much like with ignition timing (spark advance) they are what they are. In regards to the later, the actual number is not as important as finding the right point for a given engine. Why worry for example that a Chrysler "hemi" needs less spark advance that a Chevrolet "wedge" combustion chamber? The number in and of itself is not important and it is much the same with rod to stroke ratios. Unless you want to completely redesign the engine (including your block deck height etc.) leave your rod lengths alone. Let's not forget after all, most of us are not racing at the Indy 500 but rather are hot rodding stock blocks.
Only professional engine builders who have exhausted every other possible avenue of performance should ever consider a rod length change and even they should exercise care so as not to get caught up in the hype.
This is a page I saved from another website that has gone.
I did not write these articles
The sketch shows a piston moving down a cylinder bore as a consequence of combustion pressure.
This pressure is converted to a force on the piston.
(It is interesting to note that for a given combustion pressure, a bigger bore will give rise to a larger force on the piston).
The piston in turns pushes on the rod, and this force is subsequently used to create a torque on the crank, causing it to rotate. Thus the burning of fuel and air is converted to mechanical energy that can be used to propel an automobile down the road (or
The length of the rod is depicted as L1 in the figure. Similarly the length of the crank arm is denoted by L2.
Observe that L2 is not equal to the stroke. The stroke is in fact twice L2. stroke = 2 x (crank arm length)
The ratio of the rod length to stroke is called "rod ratio" and is a useful term to quantify the kinematics (relative motion) of the piston as the crank completes a cycle.
Rod ratio can also be used to quantify the dynamics of piston motion (the relative forces) but that is for another article.
The equation for the rod ratio is as follows: Rod Ratio = rod length / stroke.
The total distance that the piston moves down the bore is solely determined by the stroke of the crank.
But both the speed, and the acceleration of the piston are dictated by the rod ratio. The piston speed and acceleration can have numerous effects on the performance of an engine.
The velocity of the piston (it's speed) can be important in determining how the intake charge is pulled through the ports and past the valves.
A fast moving piston will pull harder on the ports, creating a larger Delta-P to "suck" air into the cylinder on the intake stroke. Here one might think of correlating the point of maximum piston speed to the point of maximum valve lift for example.
The acceleration of the piston on the other hand, leads to forces on the rod and main bearings, as well as on the wrist pin. These forces put a limit on the rpm's that the bottom end of the engine can reliably withstand.
The rod ratio also determines the "dwell-time" of the piston at top-dead-center (TDC) during combustion.
This means that the position of the piston relative to the point at which maximum combustion pressure occurs can be altered through changes in rod ratio.
This could be used to try to correlate the point of maximum combustion pressure to the point at which the piston has the greatest mechanical advantage on the crank for instance.
In fact, the very nature of the combustion process can be affected by the position of the piston, and how long it dwells at TDC.
These are all interesting topics, but in this article we restrict ourselves to an investigation of piston displacement and velocity, and we concentrate mostly on piston acceleration.
Once the piston displacement from TDC is known, it is a relatively simple matter to determine the piston velocity and the acceleration.
This was done via an Excel spreadsheet which generates curves of piston position, velocity and acceleration as a function of
the crank angular rotation.
A case with all three curves shown together is presented in the following link. Note that the piston acceleration was divided
by 1000 to keep it on the same scale as the other two curves.
The situation modeled here is an Evo III spec S14 engine with 87 mm stroke and a 144.25 mm long rod:
Observe that the velocity and acceleration curves are not perfect sinusoids.
They approach being perfect sinusoids as the rod is lengthened (increasing the rod ratio).
Also, the maximum piston velocity occurs well before 90° ATDC.
Furthermore, the maximum piston acceleration occurs at TDC, being roughly twice as large here as compared
But the peak piston acceleration at TDC occurs very briefly, while near BDC the piston is accelerated upwards at a relatively constant rate for almost 70 degrees of crank rotation.
This curve was constructed for 8000 rpm.
Clearly at a lower rpm the velocity and accelerations would be smaller, while the displacement would remain the same.
So, on to the question of interest: Can we use a longer rod to decrease piston acceleration, and thereby
build a bottom end capable of reliably sustaining higher rpms?
This series of curves shows that a longer rod reduces the maximum piston acceleration.
An infinitely long rod (approximated here as one that is 10 meters long) will reduce the peak acceleration by 23% (relative to a factory Evo III configuration).
But that's as low as the acceleration can be made to go with an 87 mm stroke at 8000 rpm.
As the rod gets shorter, on the other hand, the max. piston acceleration is increased, but only at TDC.
At BDC, the piston acceleration is actually reduced by a shorter rod (at least initially).
The piston acceleration curve also begins to form a characteristic "double-hump" shape. If one were to keep making the rod shorter until it was only as long as the crank arm radius (a shorter rod than this would prevent the crank from completing a rotation), then the piston essentially would come to a "sudden" stop at 90° ATDC and it would "suddenly" start moving upwards again at 90° BTDC.
These sudden stops and starts lead to infinite acceleration at 90° after and before TDC, and this is what the double-hump is starting to show.
Of course this is all pure theory, as in practice the piston and rod consume space which makes the previous example impossible to achieve.
But looking at the theoretical limits of an engineering problem is always instructive.
Now a seasoned engine builder might consider trying to package a longer rod into the existing cylinder block.
The reasons for wanting to try this can vary, and one of them might be to try and reduce the max. piston acceleration in an attempt to allow the bottom end to safely maintain higher rpms.
So let's say we want to try this on an S14 engine with an Evo III crank.
If we work real hard at squeezing the ring pack together, possibly pushing the wrist pin up past the oil scraper ring, and we
reduce the OD of the wrist pin to a minimum, then we might just be able to wedge in a 1 cm longer rod. Having accomplished this we could be quite proud of ourselves in having built an S14 capable of higher rpms due to the reduced max. piston acceleration.
But how much has the max. piston acceleration really been brought down? This is easy to determine with our spreadsheet.
The curve is blown up to concentrate on the region near TDC (0° crank rotation) in order to better see the change in max. piston acceleration.
And the answer is somewhat discouraging.
The acceleration is only reduced by roughly 1.5% after all our efforts to lengthen the rod.
Now this result should not be considered inconsequential.
For example, if the previous redline limit for bottom end integrity had been, say 8200 rpm, then it is now raised to 8323 rpm. That's something you can hang your hat on.
But depending on your application, it may or may not be worth the effort required.
Remember too that, as previously mentioned, there are additional reasons why one might want to try a different rod ratio.
The piston kinematics spreadsheet is fun to play around with.
Should the reader desire to perform further experimentation, the spreadsheet can be downloaded via the following
link: 4500 feet per minute for a well built (internally balanced, 4340NT crank, 4340 rods with cap bolts, steel pins,
forged pistons, and a suitable valve-train) street/strip engine is around the limit for reliability.
Production engine's are generally good up to around 3800-4000 feet per minute.
Racing engines such Nextel Cup, F1, Indy, and even sportbikes have piston speeds exceeding 4800 feet per minute and maybe