In 1964 when the 426 Hemi was introduced it was primarily a racing engine. It was on the 23rd of February when four Hemi-powered Mopars swept the Daytona 500 and finished finishing first, second third and fourth. This single event took the racing world by storm and ultimately led NASCAR to impose stricter production rules on Chrysler. As a result they would be required to produce several thousand Hemis in regular production vehicles rather than only a few blueprinted Hemi motors each production year. This is what led to street Hemi which was a slightly detuned motor found initially in 1966 B body Dodges and Plymouths. The 426 street hemi was rated at 425 Bhp but it’s been long said that these engines were under quoted during certification from the factory for insurance purposes. But just how much power did those ‘King Kong’ Hemis make by todays standards?
As well as addressing the above It is the intension of this technical article to cover a few further points. To go further we would like to introduce to you the concept of a 1 dimensional fluid dynamics engine performance prediction code. They’re not a new concept and quite a few exist. Now it is important not to confuse these with budget aftermarket codes such as Desk Top Dyno, Performance Trends or Dynomation: the commercial codes are far more sophisticated and extensive in scope. They’re often referred to as 1-D cycle simulation or thermo-fluids codes. OEMs as well as racing use them to great effect characterizing prototype engines to minimize dyno testing, optimising and to gaining further insight into engines. There are commercial codes available such as GT Power and Ricardo Wave and these are the codes that OEMs usually use although some manufacturers have been known to develop their own in house software for these purposes (Lotus cars and Ford Motor company have done so in the past). We don’t wish to get bogged down on the academic technical details of the ins ands outs of 1 D code. We can summarise by saying that our code is our own, a spin off written from university days and proven to be extremely accurate time and time again when compared back to back with measured dyno data. The code simulates unsteady, compressible gas flow through the induction and exhaust systems of multi-cylinder spark ignition engines. It solves for “quasi-3D” of the Navier Stokes equation using an explicit ,second order , finite difference numerical solution technique solving for mass, energy and momentum. A well built model extensively correlated to measured engine dyno data usually falls within 3% of the data. Once this model is established the effects of cam profile or changes in exhaust are quite easily and accurately reflected. In addition to our 1 D code we also had access to an engine dyno at Rennsport Systems and an original 1965 engineering technical development report by Chrysler engineering complete with extensive dyno data from their Highland park facility. It was upon this solid foundation (along with our own extensive empirical engine test data database) that we built our simulation model. It’s fair to say that this is probably one of the first applications of extensive 1D modeling to such a venerable engine.
What an engine is rated at is highly dependent upon its level of dress and the conditions to which it was tested. Before we go much further, it’s important to review the various old Chrysler definitions that correspond to their engines level of dress and testing conditions.
Cold Bare Gross: No air cleaner, released carburetor mixture, intake manifold heat passage blocked off, maximum power advance (MBT) set at each speed, laboratory exhaust system installed and exhaust heat valve open.
Maximum Torque @ rpm Maximum Power @ rpm
480 ft. lb @4400 480 Bhp @ 6000
Hot bare gross: This is the same as Cold Bare Gross except intake manifold heat is used.
Maximum Torque @ rpm Maximum Power @ rpm
472 ft. lb @4000 477 Bhp @ 6400
Laboratory Gross: Air cleaner installed, fixed jet carburetors (AFB- 4139 front, 4140 rear), Maximum power spark advance (MBT) set at each speed, crankcase ventilation valve operating, normal heat on the intake manifold (type of heat-exhaust), exhaust heat valve open, and laboratory exhaust system installed
Maximum Torque @ rpm Maximum Power @ rpm
474 ft. lb @4400 474 Bhp @ 6000
Full Net: Air cleaner installed, fixed jet carburetors (AFB 4139 front and AFB 4140 rear) normal heat on the intake manifold (type of heat exhaust), exhaust heat valve open, automatic spark advance set manually at each speed (#PF 1058 vacuum curve, #PF 4439 governor curve, set 12.5 degrees BTDC), car cooling fan (fixed mechanical) and exhaust system installed.
Maximum Torque @ rpm Maximum Power @ rpm
416 ft. lb @3600 390 Bhp @ 5600
So where do we go from here? It’s important NOT to take the dyno figures (even engine dynos) automatically as gospel without question: even methodically calibrated dynos show variance and several OEMs use this to their advantage (notice we didn’t even touch upon the minefield of after market dynos and chassis dynojets). Usually dynos are still very valuable when evaluating changes against an established baseline. The same level of healthy cautiousness should be used when reviewing the simulation data where the results are highly dependent on the assumptions and the code itself employs certain assumptions in its theory. This level of healthy pragmatism combined with an empirical database of what is realistic and what falls outside what is plausible will allow us to really get a good feel and understanding of the engine. We correlate the 1 D code model, fastidiously going over every engine system on a component by component basis. It’s a relatively easy task to get the simulation model to match a fixed set of dyno data. What is more useful is to see if the model can track changes in cam profiles/timing, exhaust system and intake geometry and still predict faithfully. The first task is to match the simulation to the measured engines airflow or volumetric efficiency curves (VE), An engines VE curve looks very much like the torque curve shape and various parts of the engine effect that curve depending on the engine rpm and load.
The figure below shows just how well a well correlated 1-D cycle sim model can capture cam timing changes in volumetric efficiency. The magnitude of the trends of these changes are captured very accurately. On the dyno the 276 period Hemi cam was timed at a maximum opening point of 104, 108 and 113 degrees timing ATDC.
If this was a project with an OEM we would outfit the subject engine with a network of pressure tappings, some static some water cooled kistler Real time dynamic along with in cylinder pressure transducers but these are not available to us for this project unfortunately. The next figure shows a simulation model matched to intake , combustion chamber and exhaust pressure tappings showing capturing of the tuning waves at various engine speeds for a 4.6 litre V8 to illustrate just how accurate a well tuned model can capture the real engine. This level of depth while useful is perhaps not required for this investigation though.
The most influential part of the intake system of a naturally aspirated engine are the intake ports. The ports of the Gen 2 426 Hemi are its entire reason of being. Based on the Chrysler RB big block base engine the Hemi spherical cylinder head allowed not only larger valves to be accommodated in a cross flow arrangement, the angled arrangement of the valves meant that the port flows were not as sensitive to being shrouded by the bore or combustion chamber walls.
In addition to outright flow of the ports it is important to consider swirl motion, tumble motion as well as mean gas velocity. One thing we’re going to get straight is the dropping this aftermarket practice of using CFM figures at 28 inches of water to classify absolutely every cylinder head under the sun. This site will continually focus on engineering related topics closer to an OEM perspective following their typical methodologies and practices. The biggest issues with using CFM at 28 inches of water is that it only is an absolute measure of restriction and doesn’t show how efficient the port flows for its design. We will use flow coefficient relative to bore size- which we call ‘Alpha K’ and also flow coefficient relative to the valve inner seat diameter itself (Fc). Different OEMs prefer to use different conventions: Lotus cars like to use a reference diameter of the throat diameter of the port while Cosworth have been known to use both the valve outer head diameter and the inner seat diameter while Jaguar and Ford Motor company also use inner seat diameter. By far using the inner seat diameter is the most common within the industry. Alpha K is a measure of how well the port is feeding the cylinder and is usually quoted as a percentage. Flow Coefficient or Fc is typically expressed as a fraction and is a measure of how efficient a particular port flows. A good Alpha K number for a typical 4 valve chamber at a peak valve Lift is anything above 16%. An exceptional number is 22% or above again at peak valve lift. In other words an 18% Alpha K at near peak valve lift means your cylinder is being fed well. Anything above 16% Alpha K is rather exceptional for an old school 2 valve style port. For Fc at 0.3 L/D (valve lift over diameter ratio) a good efficient port will flow above 0.6, while a well optimized port may achieve 0.68 to 0.72 at 0.3 L/D. A good way to illustrate and understand this is to imagine a good efficient port that flows 0.72 Fc at 0.3 L/D but in the cylinder when Alpha K is calculated only a modest 11% is achieved. What this would say to me is that there is very little to be gained by porting anymore as the ports are flowing very efficiently but the ports along with the valve heads sizes and associated dimensions are likely too small for the engine cylinder size or bore. Ocassionally the reverse can also be true.
The first figure shows a flow comparison of various 2 valve engines over the past 40 years that shows Flow Coefficient plotted against L/D ratio (Lift over diameter ratio). Because both axis are normalized it is a convenient and fair way to compare different engine sizes and styles. The broken black line represents the CD=1 line and is drawn for 45 degree valve seats. It represents the upper limit at low lift flows (when you’re constricted by the annulus of the valve) and under normal circumstances you shouldn’t be able to flow higher than that. Judging against the CD=1 line is a good sanity check to see if the Flow bench or Computational Fluid Dynamics (CFD) data is trust worthy. If your flow is some what below the CD=1 line at low lifts its most likely due to flow shrouding from either the cylinder bore itself or possibly the combustion chamber.
First comment to be made is that the 1969 Dodge “906” cylinder head 440 cu in (RB big block) port flow coefficients are surprisingly similar to the 1997 LS1. Now bear in mind that this is port flow coefficient so this means that in terms of flow efficiency they’re similar. The Dodge has larger valves (52.83mm vs 51.23) so probably flows slightly higher outright (CFM) flow numbers than the small block LS1 not surprisingly. What is surprising is that a 1969 wedge layout port could be similar to a port that’s almost 30 years more modern. You could mention in cylinder motion and how this is important and needs to be factored in. In fact straight after this part of the article we do just that. The next observation is that it’s not difficult to see why back in 1965 the Gen 2 426 Hemi really did rule the roost. Its port flows from stock castings were head and shoulders above other stock castings and compare favourably with the Chevy LS6 from 2002 (which is a wedge design so admittedly has a harder job to flow as well). The 1987 2 valve BMW M20 is included there as this port is considered a benchmark in terms of flow and motion compromise: This cylinder head is an evolution of Apfelbecks work and his legendary ‘DreiKugelWirbel’ or ‘3 lobe vortex’ Chamber he outlined in the 1960s. This port/chamber layout is a shallow angled hemi (44 degree total included angle) with a low heat loss compact chamber , good squish and great in cylinder motion as we shall soon see. I’ve also included the flow performance of the 2006 Gen 3 Chrysler Hemi engine. In terms of outright flow characteristics the flow follows the CD=1 line well at low lifts and then increases to an unpresidented outright flow figure.
To assess the flow capability of a cylinder head fully the previous plot of Fc vs L/D should always be assessed along side Alpha K vs Valve lift. Here we can see that the Dodge RB heads and Chevy LS1 flow characteristics diverge when assessed relative to the bore diameter. This would seem to indicate that the 52.83mm Dodge valves are rather modestly sized for the bore size of 110mm. When viewed in context of the modest 84mm bore size of the BMW M20 this little port flows very well and compares favourably with the LS1 Chevy engine. Finally we can see that the Gen 2 426 Hemi flows very well compared to the other modern engines only behind the late 2006 Gen 3 Hemi and the Porsche 911 (993) cylinder head- a head which I included as it was considered a modern production benchmark for a 2 valve Hemi style port/chamber in terms of outright flow until recently. The 2006 Gen 3 Hemi has outstanding outright flow for its 99mm bore size that would be very good even for a 4 valve figure. It doesn’t fair so well for a modern design in terms of in cylinder motion as we shall touch upon next.
A high Alpha K value will have a direct bearing on specific output of an engine. This is because Alpha K will be a high value if outright flow is high and the bore diameter is smaller- and having high flow for a modest bore size is an ideal ingredient for high BMEP and high specific out put (a small bore is better in terms of thermal losses and smaller flame paths for better knock limit). The relationship of the valve head diameter to the bore diameter will have a direct bearing on the Alpha K value. The next plot shows a snapshot figure of Alpha K at a fixed L/D of 0.2. It is based on an empirical data base of several 2 valve engines and shows how our comparator engines compare with others (the grey shaded area).
The straight line represents the upper limit of outright Alpha K possible at 0.2 L/D from our empirical database. Engines below that straight diametric line COULD be using the energy available for outright flow for in cylinder motion. Or put another way, when below that line there is energy potential in the flow to be used for knock mitigation. This doesn’t mean that just because you’re below that line the energy normally available for flow IS used for in cylinder motion- just potential. It does mean that when you’re at the limit line you are very likely NOT making much in cylinder motion however. The graph shows that the Porsche 993 (M64) cylinder head as well as the Gen 3 2006 hemi represent the upper limit for outright flow for their Dv/Dcyl (Valve diameter to bore size) . Unfortunately these engines do not have a lot of in cylinder charge motion. The 1960s Gen 2 426 Hemi could flow better considering its Dv/Dcyl however , as subsequent plots will show, this engine doesn’t have much in cylinder charge motion either.
In cylinder motion can be resolved or classified as either tumble motion or swirl motion (in reality most port/chamber geometry) will tend to produce a combination of both to greater or lesser degrees. In cylinder tumble motion will typically tend to speed up burn near TDC and make an engine more resistant to knock and if optimized correctly can also give the engine ability to run stable lean combustion . An example of a high tumble 4 valve engine that is capable of running very lean is the Honda S2000. The plot shows a snap shot of Alpha K for 0.2 L/D again, versus tumble intensity as measured on our flow bench. The grey shaded zone represents the empirical data we have for several 2 valve engines. The curved broken boundary curves (in red, cyan and green) show valve angles of the cylinder head design as typically a deeper angled combustion chamber will show a better flow/tumble compromise (tumble motion at TDC typically doesn’t decay as much with a deeper chamber). However too deep a chamber and you’ll get excessive heat loss and your piston will need to be domed to attain the target design compression ratio will exacerbate the problem. It’s all a case of balance. The BMW M20 head produces the highest tumble intensity even though it still flows better than poor tumbling heads like the 1969 RB 440 Dodge and doesn’t flow a lot worse than the 1997 Chevy LS1 head. When viewed in context of the Dodge Gen 2 426 Hemi engines HUGE included valve angle of 58.5 degrees the engine produces very poor tumble and reasonable outright flow – although back in the 1960s it is unlikely that tumble motion and good combustion burn rate was a priority or even fully understood. The Porsche 993 head produces better tumble and good flow considering the size of valve in the bore size. Finally with the 2006 Gen 3 Hemi (34.5 total included valve angle) it’s obvious that the priority of the engine design was outright flow over in cylinder tumble motion. The twin plug arrangement of the new Hemi will no doubt help with knock mitigation and stable combustion (by shortening flame paths) and this is the reason why the ports were prioritized towards outright flow (See SAE paper 2002-01-2815). Just for comparison and to show how far a well developed 4 valve engine has come I’ve plotted the 4 valve Honda S2000 cylinder head figures on the same 2 valve plot space. The Flow-tumble motion compromise is in a totally different range.
In cylinder swirl motion is the other component widely considered. Typically this doesn’t benefit knock limit mitigation much but if done correctly can speed up burn rates (late in the cycle) and help with engine lean running capability. Engine lean running capability will benefit the engines ability for stable combustion with higher EGR rates also. At 0.25 L/D the BMW M20 head produces very respectable flow and a swirl ratio that is on a par with the LS1, Dodge 440 even the VW Polo 1.2. The Honda Jazz 2 valve makes a very high swirl ratio at the expense of flow. The Porsche M64 and Gen 2 426 Hemi cylinder heads produce the highest flow out of the bunch presented and the ports were not designed to impart swirl motion.
In summary the Gen 3 Hemi flows outstandingly well but with poor in cylinder motion- however its twin plug design will no doubt help in this area. The Gen 2 426 Hemi flows respectably even by todays standards compared to an GM LS6 engine, although there’s certainly room for improvement in these heads using modern techniques. The priority with the 426 Hemi was obviously not in cylinder charge motion. The stock 1960s Dodge RB ‘906’ cylinder head is surprisingly similar in Flow coefficient to the LS1 Chevy small block (one wonders how the previous LT1 engines faired) although the RB big block could benefit from larger valves. The 1980s 2 valve BMW M20 cylinder head represent a good compromise of both outright flow, in cylinder swirl and tumble motion. This probably explains why this engine produced the highest specific output for a NA 2 valve engine in its day with good resistance to knock and lean running ability. Not a lot of porting work would be required to significantly improve these heads in all areas. The Porsche 2 valve 993 head is not dissimilar to the Gen 3 Dodge Hemi. The priority wasn’t placed on in cylinder charge motion but again the design has a twin spark plug arrangement. We wouldn’t be surprised if the 993 Porsche cylinder heads were used as a benchmark by Chrysler to develop their Gen 3 Hemi heads.
If port flow was the only criteria to judge good ports, ports would end up being drain pipe sized (much like the less technically accomplished Detroit manufacturers tried in the 1960s with Tunnel ports and ‘Ram Air’ heads) presenting absolutely minimal restriction. In reality there is a compromise between port velocity and port flow. This fact is actually lost TODAY amongst some of the Korean manufacturers. Too large a port with slow port velocities and the in cylinder burn will be slow due to poor motion and slow momentum of incoming charge. This is doubly important on a carbureted engine where the flow is essentially wet with fuel droplets in suspension. To asses port velocities the port is split into the entry portion, mid portion (or trouser leg in the case of a 4 valve style head) and the throat. The guidelines used are calculated as mean gas velocity at peak power speed. Good design practice guidelines for a high performance engine are 100-110 metres per second , 90 to a 100 m/s and 70 to 80 m/s at the Entry/mid port/ and port throat respectively. These are rough guidelines and good engines can still stray outside these slightly. Engines with fully machined ports good flowing ports will tend to target higher velocities. The BMW S54 M3 engine for instance achieves 110./119/94.3 m/s at its 7400 rpm peak power speed.
With the design peak power speed of the Dodge 440 six pak engine of 4700 rpm and 6000 rpm for the 426 Hemi engine you can see that the port velocities of the two top big block Mopar muscle car engines are right in line with guidelines. It’s obvious that Chrysler engineers in the 1960s knew what they were doing and , as indicated, this wasn’t necessarily the case with their period peers.
Carburation and Intake Losses
Intake losses on cars with carburetters is a complicated matter and always a case of compromise. Not too big because you’ll lose driveability. Not too small or the carb becomes a major bottleneck. For most hot, dual-purpose cars, pulling about 0.9 to1.0-inch-Hg manifold vacuum at WOT, max rpm on the dyno is a good target. Now I’m quoting this figure using the Hot rodders norm of having no air cleaner or induction trunking fitted with the venturis of the carbs drawing straight from atmosphere. This is what’s known in the industry as Gross power figures and was the norm even for OEMS homologating back in the late sixties. Once you put on the period air cleaner and induction system you’re likely to be in the 1.15 to 1.35 inches of HG region. It’s quite an easy task these days however to source a less restrictive (but noisier) air cleaner induction set up for carbs that will bring this down to virtually no difference from the target 0.9-1 figure. These are all guidelines from racing experience but there are no guarantees. It’s impossible to know for sure how a carb will perform in a vehicle based on how it did on an engine dyno test in the lab. Every combo behaves differently when you’re actually out on the road with a lot of transient accelerations and various road gradients. What works in a featherweight car with a manual trans and 5-series rear-gears may not be ideal in a 3,700-pound car with a mild converter and 3.23:1 final drive.
Carburetor sizes usually quoted in inches refer to the diameter of the barrel. The Dual quad 426 Hemi tested in 1965 used a Carter AFB 4139 at the front and a AFB 4140 at the rear. These were chosen to minimize the cylinder to cylinder Air fuel ratio variance as much as possible. The barrel diameter or throttle bore diameter of the primaries were 1-7/16 and for the secondaries it was 1-11/16. This fed into a dual plane intake manifold which connected together equal phase (360 degrees apart) fired cylinders in sympathy with the firing order (which was common practice back then). What is more important than the barrel diameter and what was common practice to quote was steady state pressure drop across the carburetor as measured on a flow bench. Some manufacturers, such as Holley, would have this information readily available. It was more difficult to get this data from Carter but not impossible. We found out that the Carter AFBs used on the Hemi were rated at 570 CFM each. Flow capacity gives more important technical information than just the barrel diameter as for a given barrel diameter the venturi size can be different or other subtle but significant details that effect the 3 D flow and therefore the restrictiveness of the carburetor. The next thing to note is that most manufacturers rated carburetors in terms of DRY flow which is quite different to wet flow. Dry steady state flow is about 8% higher than wet. Wet flow is what is closer to how an engine is run. 4 barrel carburetors were typically rated at 1.5 inches of mercury pressure drop. We built a mock up of a steady state rig in simulation to mimmick the steady state pressure drop of the original carburetor test conditions. The restriction of the modeled AFB barrels and venturis were adjusted until the Carter ratings were obtained and this calibrated Carb model was placed into the engine simulation. The match to the 1.25 inches of depression of the actual running Hemi engine fell to within 10-15% even when Dry flow rating to wet flow was taken into account (the real engine carburetion was more restrictive). We went with the measured running engine depression. An engine runs dynamic pulsing flow that continually varies with time as distinct to the steady state rigs continuous pressure drop.
There could be a number of reasons for the less than perfect match up of restrictiveness of the carburetor rig data to the engine measured depression including unknown entry and exit conditions of the steady state test rig, subtle variations of the exit conditions of the mounted carb on the engine and unusual 3 Dimensional pulsing effects on the running engine that effected results.
Engine Camshaft profiles
A profile on profile comparison is shown in figure ##. The 426 Hemi solid tappet cam profile is shown up against to other notable pushrod engines: The Buick derived Rover V8 and a late Gen 3 5.7 Hemi. The first thing that is striking is the length and height of the ramps of the Gen 2 426 Hemi. This is in part due to the fact that we’re comparing a flat tappet solid lifter cam with hydraulic lifter cams (Rover V8 is a flat tappet hydraulic and the Gen 3 Hemi is a roller hydraulic). Despite this the height and length (read mild ramp rates) of the ramps seem excessive by modern day standards and will effect engines in cylinder residual level (blow back of exhaust gases back into the cylinder) at light loads and low speeds.
Closer examination of the 3 profiles shows that the Gen 2 426 Hemi cam flank accelerations are quite modest considering that it is a solid lifter profile (solid lifter profiles are usually able to have higher peak flank acceleration). It is lower than the Gen 3 Chrysler Hemi (which is a roller hydraulic) and the Rover V8 hydraulic flat tappet cam which is surprising. This is probably due to the age of the design and possibly the mass of the valve system (Tappet, pushrod, rocker arm, valve itself etc) of the Gen 2 Hemi. A glance at the profile velocities shows than the gen 2 426 Hemi with it’s wide 0.903 in diameter tappets allows highest peak velocities for a flat tappet cam compared to the Rover V8 (with its 0.842 in tappet diameter) but predictably the Chrysler Gen 3 with its roller tappets gives the most freedom to the cam designer by allowing very high peak velocities. Although the total cam period or duration of the Rover V8 and Chrysler Gen 3 are of a similar order as are the peak flank accelerations the Chrysler Gen 3 gets its considerably higher peak valve lift from higher acceleration across the cam nose. The Rover V8 and Gen 2 426 Hemi seem to have acceleration over the cam nose of a similar order.
In summary the Gen 2 426 Hemi profiles have long and tall ramps which are not good for performance and can probably be improved upon by adopting modern cam design practices. Similarly the nose and flank acceleration could also probably be further optimized with modern thinking. The long tall cam ramps will effect in cylinder residual levels and increase overlap thus adversely effecting cylinder to cylinder distribution (a particular problem on a high performance carbureted V8).
Engine friction evaluation/characterization
Engine friction is an often overlooked aspect of tuning and power. What is good for reducing friction will benefit fuel economy too. In the same way that increasing compression ratio benefits both also. We will look at engine friction of the 426 Hemi on a component by component level using a combination of our vast empirical test data base, our friction scoping tool which is a model based upon Heywoods model (see paper 2003-01-0725) but honed by applying it and adapting the theory to fit our empirical database and finally doing a sanity check to see if it all tallies up with measured motored engine data recorded. There are two ways to assess engine friction.
1) Using a motoring dyno
2) Using data from a fired dyno test from an engine outfitted with in cylinder cylinder pressure transducers (where it can be calculated from IMEP-BMEP).
Both methods have their pros and cons (it goes without saying that back in the 1960s when the 426 Hemi was conceived there were no in cylinder pressure transducers to get indicated mean effective pressure data). With the motoring test there is no way to account for increased local temperatures due to combustion (the colder wall temps during motoring will tend to give worse reciprocating friction than reality) and because there is no combustion the gas load on the bearings is a lot lower than a fired engine. These effects are somewhat self canceling and the numbers from motoring are not only good for relative differences but also a good first stab for absolute friction assessment. The in cylinder indicated transducer method is prone to errors because FMEP is small compared to IMEP and BMEP so any measurement error is amplified. The in cylinder transducer method is also very intrusive and involves machining and adapting cylinder heads to accept in cylinder transducers (more recently there are spark plug adaptor transducers but these are costly and hard to come by).
First of all we can see the projected rotational friction of the 426 Hemi crankshaft main bearings. The figure shows this up against favourable engines (green dotted curve), middle ground engines (yellow) and unfavourable engines (red dotted curve). This is perhaps a surprising result as it shows that the 426 fairs very well even compared our predominantly modern engine data base. The 426 Hemis main bearing size is 69.85mm which is similar to contemporary BMW V8 or older Porsche 928 engine, both of which are much smaller in capacity (7 litres versus under 5 litres) . However bearings are sized with peak cylinder pressure of the engine in mind and this is where the Hemi is mild compared to a European V8. I would even venture to say that the bearing could be made smaller especially when you consider the WIDTH of the Hemi bearings.
Reciprocating friction is the friction caused by the contact of the piston rings to the bore, the piston skirts to the bore and is also a function of the stroke length and engine speed or ‘mean piston speed’. In terms of piston design modern engines have come a long way. Skirts are a fraction of length they used to be and they are often of a slipper skirt design where they only exist on the thrust and opposite sides. Ring lands have come down dramatically and ring widths have come down, while rod length to stroke length ratios have increased. Rod length to stroke length ratio determines the angularity with which the thrust side of the piston is driven into the bore. As compression heights of pistons have come down rod length have been able to go up. Rod length to stroke ratio is an area where the Hemi and old RB big block excels compared to 1960s Big blocks of the time (Chevy Big Blocks deck height was only 9.8 inches vs the Mopars 10.72) .With all of this in mind it comes as little surprise that in terms of reciprocating friction the 426 Hemi is only slightly better than the mid band of benchmark contemporary engines out there. Typically when these engines get modified the pistons used have minimal slipper skirt design pistons with longer rods and small compression height. It is also common practice to grind down the stock con rod journal diameter of 60.2 mm down to the Chevy diameter of 55.9 mm which also helps bring down the reciprocating friction.
The valvetrain friction figure shows different boundaries of friction of various types of valvetrain from measured motoring data from our data base. Of all the types the highest friction is typically exhibited by poorly optimized sliding contact OHC valvetrain with hydraulic compensators (as part of the reciprocating mass). The lowest friction is usually exhibited by lightweight valvetrain outfitted with good roller design. The Cam-in-block design or pushrod is a venerable design but fairs reasonably well due to a number of reasons:
- For a given V engine there is usually only one cam shaft and associated number of bearings rather than having 2 or 4 with an OHC layout.
- The camshaft is located in a cooler part of the engine which helps the regime of lubrication (an OHC typically falls under mixed lubrication regime (which is the highest in terms of friction) while the cam in block will tend towards elastrhydrodynamic regime which has a lower coefficient of friction)
Out of the pushrod engines the graph shows the Gen 3 Hemi valvetrain friction (which doesn’t fair that well) with the 426 Gen 2 hemi and a Big Block wedge engine fitted with needle bearing rockers that have roller tips (but still flat tappet). What counts against the pushrod layout is the sheer mass of the valve train system and the number of rubbing surfaces. Without a doubt the way to better a pushrod layout is to go to roller tappets, rocker arms with needle bearings and rollers and a solid system with no hydraulic compensators which helps lower friction a lot. Out of the pushrod systems shown the 426 Hemi system fairs best mainly because it has solid lifters (and a relatively tame profiles).
The overall plot shows how well the empirical friction scoping tool compares against measured motored engine data (where pumping losses have been subtracted). The tool allows better understanding of where to optimize further when trying to raise the engines performance.
During the early to mid sixties directly driven fixed fans were commonly used to cool the big barges across the USA. The early Mopar big blocks were no exception. The engine used to homologate the Hemi in 1965 used a 4 blade fixed driven fan. This was tested in before and after testing (where dyno cell cooling was used to keep the engine cool) and the losses of the fan derived. It’s important to note that even with only a 4 blade fan the losses at full power are quite significant. The figure shows over 0.3 bar FMEP at full power. This amounts to 15 to 19 Bhp losses at peak power. It’s important to note that the Cold bare gross number of 480 Bhp doesn’t include this fan. The Lab Net number does include the direct driven fan. The later hemis and indeed the majority of engines post the 1970s started using viscous coupling clutched fans that slip above a certain engine speed and when the engine is receiving sufficient cooling (high vehicle velocities) and therefore reducing the frictional losses.
Engine combustion and ignition values
The final significant ingredient of the simulation model that is to match measured dyno data is combustion or some reasonable approximation that comes close to what was measured on the real engine. The full load ignition advance is shown in table.
These are optimized values that were obtained on our dyno at Rennsport and at the original test in 1965 in Chryslers Highland park facility. Optimised means MBT timing or where the ignition timing is swept until the best value of torque is obtained. Any more or less ignition and the torque falls off. When engine performance development engineers look at combustion of an engine they are concerned with the burn duration. This is typically split up into ignition delay and the main burn duration. The ignition delay is the time between when the spark is initiated and the spark nucleus is formed to when the laminar flame front gets underway to start burning across the combustion chamber. The ignition delay is hard to measure precisely so is usually quoted as 0-5% burn duration (or the time from ignition spark to when 5% of the in cylinder charge is burned). It’s important to bear in mind that the ignition delay will tend to increase with increasing engine speed even though the time taken is similar in milliseconds the time expended in crank angle becomes more. Similarly burn duration is usually quoted as 10-90% burn duration. The apparatus used to measure this are in cylinder pressure tranducers (as touched upon previously)-which were not available during the development of the 60s Hemi (or for our test work at Rennsport). The relationship between Ignition advance and burn duration and ignition advance and ignition delay is not a clear cut one and is effected by many parameters but given enough empirical measurement data clear enough patterns are exhibited to move forward. The next two figures show some of our empirical measured burn data hand picked for having ignition advance values of a similar value and similarly optimized for MBT timing. The burn duration and ignition delay values we picked to populate the model ended up being at the slower side of the bands as we already know that the Gen 2 426 Hemi engines ports and chamber offer little motion and therefore a quiescent burn (therefore a slow ignition delay and slower burn).