Thanks for the link

 

I didn't have trig in school, they did away with it by the time I would have taken it, so I'm thinking maybe that is why these formulas don't make a whole lot of sense to me.
The formulas that just don't make sense to me are:

P = R cos(q) + (L2 - R2 sin (q)2) 1/2
Se = R + L + R cos(q) + (L2 - R2 sin (q)2) 1/2
CP = CRe¹·² (A) - A

In the first one, why is it L2, normally it is 2 times L or 2L or was that supposed to be L² or L squared? As can be seen in their formula here: Ve = B² (ES) .7854 they have a square displayed properly so not sure why the L2 and R2 if to be squared wouldn't be displayed properly as L² and R² respectively.

Then in that same formula why is it 1/2 at the end... Wouldn't you just do /2 or divided by 2?

R times cos of (q) plus parenthesis L2 minus R2 times sin of (q) times two or is it squared (q)² end parenthesis ?times? one half?

The other one that I've never seen in my life is "CRe¹·²" I didn't know we could do powers in decimals. So that is CRe (compression ratio effective) to the power of 1.2?

I know I'm rusty on my math but I didn't think it was that bad... Are the formulas taken from here straight forward to everyone else or is there some issues with how they are written?

I've been scouring Google for dynamic compression formula and dynamic compression equation and it's so very convoluted out there. And there is a lot of differing opinions on what dynamic compression is. So say it's effective compression ratio, some say it's the smaller ratio that is done once the intake valve closes (this is my understanding)... And next to no one displays a straight formula and many say the calculators are all different and likely incorrect as they all give different answers to the same base numbers.

This is one hell of a can of worms I've opened.

Oh, and I've bounced all over that ftlracing website (as it looks to be the most complete with the math and formulas etc) and there is not any sort of contact information anywhere to get clarity from the source. :/

I also found this where Scott Tucker posted up a worksheet that he called effective compression but looks to me to be dynamic compression. It might help clarify the formulas:

https://honda-tech.com/all-motor-nat...ratio-2683949/

I think we may need to set definitions for both dynamic compression and effective compression so we are all on the same page.

Here is my take on them. For our purposes of this discussion I believe they should be defined as:

Dynamic Compression: The compression ratio of the effective stroke regardless of temperature, air density etc. The determining factor of dynamic compression is valve timing or better known as Intake Valve Closing (IVC). What makes it dynamic instead of static is cam timing can change the ratio by advancing or retarding the cam.

Effective Compression: This is the dynamic compression plus real time variables of temperature, boost, air density etc taken into consideration.

Edit: Okay learned something new about Math I never learned. And basically the decimal exponents are possible, used to be a slide rule or logarithmic tables or something of the like. Simply, use a scientific calculator and let it do the decimal exponents for you. Now just need to get clarity on the other formulas as they don't seem to be written very clearly using the general rules of algebra.

 

It does indeed. I'm in.

In that context the 1/2 does appear to be "divide by 2"...

L2 and R2 written as such would usually mean the second item attribute or second measurement or whatever... but that doesn't make so much sense in this context, given we are talking about rod length and 1/2 stroke...

Where is this RS article the writer is referring to?

Something from my university math is trying to get out of the fog in my brain but I can't quite put my figure on it...

 

Additionally, they are writing where P is position ATDC:

P = R cos(q) + (L2 - R2 sin (q)2) 1/2

I found another formula that calculates position location ATDC, but is written more clearly.

HT = (r + c) - (r cos (a)) - SQRT(c^2 - (r sin (a))^2)
where HT = position location ATDC, r = stroke/2, c = con rod length, a = radians of crank angle.

Let's rewrite this one using the same variables we have above: note "q" is not defined but must be crank angle as well.

P = (R + L) - (R cos (q)) - SQRT(L^2 - (R sin (q))^2)

and compare:

P = R cos(q) + (L2 - R2 sin (q)2) 1/2

It appears to me that the formula on FTLRacing is written incorrectly. Since we're not dealing with a second item or measurement - it looks like it should read:

P = R cos(q) + ((L^2 - R^2 sin(q)^2)/2), unless they also messed up the 1/2 and intended ^1/2 as the square root, which I'm thinking is probably the case, so it becomes:

P = R cos(q) + SQRT(L^2 - R^2sin(q)^2)

and compare again to the new formula I found (converting the variables for easier reading)

P = (R + L) - (R cos (q)) - SQRT(L^2 - (R sin (q))^2)

bingo. remember (R sin(q))^2) = R^2sin(q)^2

(which just "proves" that they meant R^2 instead of R2... and SQRT instead of 1/2.

Poorly written if I do say so myself. Keep in mind I haven't looked at a single compression formula before, or RS formula or anything, so forgive me if I'm missing some basic points... But hopefully this helps in some way!

I do think this will be your answer: HT = (r + c) - (r cos (a)) - SQRT(c^2 - (r sin (a))^2)

And the link I provided works you through some calculations so it should help prove if you're on the right track.

 

I think you are absolutely right. And yes Q is rod angle. The RS article is here but once again didn't do the formula writing correctly. I suspected the L2 and R2 was supposed to be L² and R² respectively or better written like you did L^2 and R^2. And the 1/2 makes sense now for the square root. ()^½

I think if we can collect and nail down all the formulas needed to evaluate a motor into one spot, it will be a useful tool for everyone. Then the challenge is getting the base data to the parts.

One thing I'm learning from all this, I stopped learning math prematurely.

Almost forgot. THANK YOU!! I've spent hours trying to locate another formula for the same thing and not having luck. I'm glad you did. Sadly, think the other formulas will be similiar. It's the only place they are all in one spot. Take them, clean them up to be accurate.... Excellent tool for all.

 

Okay now that we have the correct formulas to determine effective stroke/dynamic compression....

How do we translate the manufacturer data into Intake Valve Closing data?

Lobe center	110
Rocker Ratio: Intake / Exhaust	1.6 / 1.8
Advertised Duration (.010") Intake at the Valve: Primary / Vtec / Secondary	293 / 292 / 293
Advertised Duration (.010") Exhaust at the Valve: Primary / Vtec / Secondary	/ 302 /
Duration @ .050 Intake at the Valve: Primary / Vtec / Secondary	197 / 228 / 197
Duration @ .050 Exhaust at the Valve: Primary / Vtec / Secondary	/ 208 /
Lobe Lift Intake: Primary / Vtec / Secondary	.205 / .254 / .205
Lobe Lift Exhaust: Primary / Vtec / Secondary	/ .214 /
Gross Lift Intake: Primary / Vtec / Secondary	.328 / .406 / .328
Gross Lift Exhaust: Primary / Vtec / Secondary	/ .385 /

I also don't see the Lobe Seperation Angle provided...

The numbers in the table seem to be the typical figures given by cam makers. And I'll be the first to admit I'm a total noob here.... I can't say I understand exactly what all those numbers mean.

The only ones that I easily deciphered was Lobe Lift being the inches the lobe lifted from the base circle of the lobe.

Not sure what the difference is with gross lift.

Not sure if the numbers in Duration @ .050" is Degrees i.e exhaust duration is 208°. And if it is angle relevant, what does that mean. I understand the .050" means the point the valve is lifted .050" but that's about it.

And lastly, lobe center at 110° is that the position of the center of #1 lobe?

And just for a touch of clarity, the degrees being discussed in cams is all relevant to the crank shaft? Another words when we say 110° lobe center on the cam that is to say the crank shaft is at 110° After Top Dead Center?

I wasn't lying when I said I was a noob. But this noob is eager to learn.

 

Hm... Well, I've never looked into cam data before so I had to do some learning. (if you don't mind another noob coming along for the ride)... I'm interested to learn this.

Where did you get the manufacturer data from? any other clues on that document?

So far I've come up with:

"Lobe separation can be measured using a dial indicator and a degree wheel, but is usually calculated by dividing the sum of the intake centerline and the exhaust centerline by two.

The intake centerline is the point of highest lift on the intake lobe. It is expressed in crankshaft degrees after top dead center (ATDC). Likewise the exhaust centerline is the point of highest lift on the exhaust lobe. It is expressed in crankshaft degrees before top dead center (BTDC). The cam centerline is the point halfway between the intake and exhaust centerlines."

However, "lobe center" as stated in the manufacturer data seems vague. Do they mean cam center? If we had the intake and exhaust centerlines we would have lobe seperation.

Hopefully you don't mind my posting this here, but this helped clarify what advertised duration is for me...

"Advertised duration is the angle in crankshaft degrees that the cam follower is lifted more than a predetermined amount (the SAE standard is 0.006") off of its seat. Duration @.050" is a measurement of the movement the cam follower, in crankshaft degrees, from the point where it’s first lifted .050" off the base circle on the opening ramp side of the camshaft lobe, to the point where it ends up being .050" from the base circle on the closing ramp side of the camshaft lobe. This is the industry standard, and is a good value to use to compare cams from different manufacturers. Both are usually measured with a dial indicator and a degree wheel."

this is off Lunati's site, they have this page on "understanding camshaft specifications"

I need to do a bit more research here to catch up.

 

I don't mind at all you or anyone posting. I see this as a group effort for everyone's benefit.

I too was starting to look into cam spec definitions to make heads or tails from the provided cam specs.

The specs came directly from Crower website for the 63441Y camshaft. Once I know how to interpret their specs to IVC and other things, I was planning on doing that to at least 3 of the cams.

On another front, I stripped down the Z6 intake manifold tonight and have measured the runner area. These numbers might be off +/- .5mm but should be close enough to get a decent idea of rpm sweet spots etc.

So the runner width is 41.5mm and the height is 29.5mm. The half circle radius on each side of the runner center (the runner is not rectangular instead has rounded sides) is 10.5mm. So if we subtract the two half circles we are left with the square/rectangle portion of 20mm X 29.5mm which equals 590 square mm (mm²). Then the circle area turns out to be 346.36mm² with the formula of "area = pi * radius²"

Adding it up, the cross sectional area of the Z6 runner is 936.36mm².

I'm thinking to get a solid idea of the runner length of just the IM, I need to get some clay that I can roll and thread through the runner. Take length measurements from the roof of the runner, then the floor of the runner and if they differ, take the average so should represent the midline of the runner.

And the plenum diameter is 56mm for pretty much the full plenum. It does do a taper at the end just barely past runner #1 (#1 is for Cylinder #1).

 

As far as I can tell, the Duration is Crankshaft degrees.

I'm going to guess Lobe Center is for the First Cylinder Intake Lobe

And I guess you also assume a symmetrical grind.

Utilizing this tidbit I found here:
It looks like it's quite manageable to discern IVC event.

Primary/Secondary is 197° @ 0.050" And realistically, lash is 0.06" for this cam cold, heated may well be 0.050" so I think it's safe to assume that will be the beginning of the opening height of the lobe.

So 197°/2=98.5°
Lobe Center is 110° ATDC
So 180°-110=70° after Intake Center Line to reach BDC
98.5°-70°=28.5° ABDC IVC.

Then for vtec 228°/2=114°
114°-70°=44°

My brain still doesn't want to adsorb this but I am glad I finally found someone out on the big **** internet who explained converting Duration with Lobe Center to the Intake Valve Closing event.

Now we can start plugging numbers in to see what the dynamic compression ratio should be.

 

Dang great find Nancy Drew! But seriously, that takes a lot of patience. It's so hard to find good info on this sort of thing, I get angry when I have to scour the internet over for research. Glad to see you're making progress

 

It's slow work and not all done by just me. Mithious saved the day on the complex formula that I spent 2 or 3 days coming up empty.

I'm hoping to piece everything together so we have all the accurate information here for everyone to use.

Here is the beginnings of the corrected formulas. This post will have them all once done. Will need some help clarifying some of the later ones.

Also of note, I am doing serious overkill on the formulas and adding every multiplication sign etc so there is absolutely zero doubt.

Also, the trick for the "²" sign is hold alt down and hit 0178 on the number pad. Someone might need it in a reply.

Basic Engine Calculations

Displacement

Displacement is the amount of volume that is displaced as the piston are moved from the top of the stroke to the bottom of the stroke.

Displacement = bore2 * .0007854 * stroke * #of cylinders

Bore and stroke in mm.

Displacement in cc, divide by 1000 to get litre(s).

Compression Ratio (CR)

The static CR is the pressure ratio between the uncompressed air fuel mixture when the stroke is at BDC (bottom dead center) and the compressed air fuel mixture when the stroke is TDC (top dead center).

CR=(π/4*b²s+Vc)/Vc

CR=(π/4 * b^2 * s + Vc)/Vc

CR=(π(b/2)²s+Vc)/Vc

CR=(π * (b/2)^2 * s + Vc)/Vc

b=cylinder bore (diameter)

s=piston stroke length

Vc=clearance volume. It is the volume of the combustion chamber (including head gasket). This is the minimum volume of the space at the end of the compression stroke, i.e. when the piston reaches top dead center (TDC). Because of the complex shape of this space, it is usually measured directly rather than calculated.

CR = (V + H)/H

Based off of the wikipedia formula above we can decipher the variables of this formula from FTLracing.com

V=cylinder volume of the block.

H is chamber volume. Chamber volume is the volume of the head per cylinder minus piston volume (above deck

height).

V= bore² * .0007854 * stroke

Chamber volume if CR is known

If Compression Ratio (CR) is known the CR equation can be re-arranged to find chamber volume (H).

H = V/(CR-1)

Rod Stroke Ratio

Three variables affect piston movement: Crank Angle, Stroke, and Rod Length.

Four variables affect piston velocity (and acceleration): angular velocity, Stroke, and Rod Length.

The mean speed of the piston is determined by RPM and stroke.

Mean Piston Speed = 2 * stroke * RPM/12 (ft/minute)

This number is simply the calculated average speed of the piston. If the piston(s) were moving in 1 direction (or in a perfect circle) at one speed this would be that speed (circular velocity). This simple calculation can be helpful in quickly comparing average piston speeds and the affects of stroke on piston velocity. Unfortunately this calculation is often misused by the uniformed and used as proof that Rod Length (Rod Stroke Ratio) does not affect piston speed. We know this is not the case.

To understand the motion of a piston it is helpful to visualize the actually problem. Below we have a diagram of a piston, rod, and crank.

P = the y axis position of the piston (blue line)

L = the rod length (black line)

R = ½ the engine stroke (yellow)

q = the crank angle

a = rod angularity

Figure 1: Piston, Rod, & Crank Diagram.

We know the stroke and rod length so for any given crank angle we can determine the piston position using the law of cosines.

P²+R²-2RPcos(q) = L² (Law of Cosines)

P^2 + R^2 * R * P * cos(q) = L^2

So P = R cos(q) + (L² - R² sin(q)²)^½

P = R * cos(q) + sqrt(L^2 - R^2 * sin(q)^2)

To find velocity and acceleration of the piston we need to convert the crank angle (q) to angular velocity (w) and time (t).

q = crank angle (degrees)

w = angular velocity

t = time

U = Revolutions Per Minute (RPM)

p = position at q

q = wt

w = 2p(RPM) (radians/minute)

for simplicity sake RPM = U

w = 2 *p * U (radians/minute)

This gives us an equation for position as:

P(t) = Rcos(2pUt) + (L² - R²sin(2pUt )²)^½

P(t) = R * cos(2 * p * U * t) + sqrt(L^2 - R^2 * sin(2 * p * U * t)^2)

To determine the position at q degrees (p) insert a time of 60s/((U)(360/q))

I think this is saying t = 60seconds/(U * (360/q))

To determine velocity and acceleration we need to derive the position formula.

V= d/(dt)(Position)

A= d/(dt)(Velocity)

V(t) = -Rsin(2pU)pU - 2R²sin(2pU)cos(2pU)pU

(L² - R²sin(2pUt)²)^½

V(t)= -Rsin(2pu) * pu - 2R² * sin(2pU) * cos(2pU) * pU/(L² - R² * sin (2pUt)²)^½

A(t) = -4 sin(2pu)p²u² - 4R⁴sin(2pU)²cos(2pU)²p²U² -   4R²cos(2pU)p²U²   +   4R²sin(2pU)p²U²

(L²-R²sin(2pUt)²)^3/2       (L²-R²sin(2pUt)²)^½   (L²-R²sin(2pUt)²)^½

A(t) = -4sin(2pu)p²u² - 4R⁴sin(2pU)²cos(2pU)²p²U²/((L²-R²sin(2pUt)²)³)^½ - 4R²cos(2pU)p²U²/(L²-R²sin(2pUt)²)^½ + 4R²sin(2pU)p²U²/(L²-R²sin(2pUt)²)^½

OK, so what does this calculus mean?

It means BOTH stroke and rod length affect velocity and acceleration of the piston.

If you increase stroke, you increase piston velocity and acceleration.
If you reduce rod length, you increase piston velocity and acceleration.

Rod stroke ratio (RS) is often used to describe an engine's piston motion. RS is just the Rod Length/Stroke. Rod stroke ratio is important because it determines wear, velocity, and acceleration of the pistons. From figure 1 we can see that RS ratio directly affects rod angularity (a) along with Rod length and Stroke (definition of RS).


Figure 2:Acceleration Rod Stroke Ratio comparison. Stroke constant at 3.5", Rod length (5" & 7")

Note: All velocity and accelerations graphs in this article were created using 8000 RPM.

Above we can see how rod length affects piston acceleration. The graph details the piston acceleration for two engines on with a 5" rod and the other with 7" rod; both with strokes of 3.5". Note the graph is plotted in g forces. The G's on the pistons is fantastic. From the graph you can see that the negative acceleration after TDC is the greatest. This explains why Rod failure often occurs at the point.


Figure 3: Piston Velocity Rod Stroke Ratio comparison. Stroke constant at 3.5", Rod length (5" & 7")

Above we can see how rod length affects piston velocity. The graph details the piston acceleration for two engines on with a 5" rod and the other with 7" rod; both with strokes of 3.5". The short rod engine (blue) achieves a higher peak velocity and achieved the velocity sooner (which affects engine breathing).

Engine wear and failure issues:
* Rod angularity affects piston side loading. Reducing the maximum rod angle (a) will reduce side loading.
* Reducing piston acceleration from and toward TCD (point of maximum acceleration) reduces tensile loading of the rod.

Engine ignition and breathing issues:
* A piston that dwells at TDC longer allows the air/fuel charge a longer time to burn. This allows less ignition timing for peak power. Less ignition timing is useful because it reduces detonation allowing (slightly) higher compression ratios.
* A piston that dwells at TDC shorter increases the speed of the exhaust gasses during the overlap period. This increases the scavenging effect at low rpm and the engine makes more torque at low RPM.
* Reducing and delaying peak piston velocity allows the intake valves more time to open more to fill the cylinder. This allows a smaller intake running volume (and plenum) and better high RPM breathing.

Effects of a longer Rod
* Less rod angularity reduces wear.
* Lower piston velocity and acceleration reduces tensile loading of the rods.
* Less ignition timing is required which resist detonation.
* Compression can be increased slightly before detonation is a problem.
* Less intake runner volume is required and high rpm breathing is improved.
* Reduces scavenging at low rpm (weaker low RPM power).
* Longer TDC dwell time. (high RPM efficency).

Effects of a shorter Rod
* Increased rod angularity increases wear.
* Increased piston velocity and acceleration increases tensile loading of the rods.
* Increases scavenging at low rpm (increased low RPM power).
* Reduced TDC dwell time. (Reduced high RPM efficiency).

Stroking an engine is when you increase the stroke of the engine at the expense of rod length. When stroking an engine the rod length is decreased because the deck height is constant.

Effects of stroking an engine
* Increased displacement.
* Increased rod angularity increases wear.
* Increased piston velocity and acceleration increases tensile loading of the rods.
* Increased scavenging at low rpm (increased low RPM power).

Stroking an engine can be a very effective way to make more power but close attention should be paid to piston velocities and accelerations. Stroking an engine is a "double whammy" on increasing the piston velocities, stroke is increased and rod length is reduced.

Figure 4 & Figure 5 below demonstrate this double whammy effect on piston acceleration and velocity. The comparison used below is between a Honda B20 block and a B16A block. Note: that the block height of the B20 is actually 7mm taller than the B16a. A B16a block with the B20 crank would have even slightly higher piston velocities.


Figure 4: Rod Stroke Ratio Acceleration, B16a (blue) vs B20 (red)

This graph clearly illustrates why engine wear is increased and rod failure is possible if rev limits are pushed to far with B20 blocks. The g forces the rod experiences are significantly more (~20% more) at the same RPM.


A lot of misinformation exists on the web about the effects of RS ratio and Rod Length's effect on piston velocity. We at FTLRacing hope you've found this article useful. Further discussion of Honda hybrid (stroker) engines can be found in the Honda Engine section at ftlracing.com

This comes straight from FTLracing and I hope I corrected the formulas correctly. The last portion is getting into calculus and hard core trigonometry which neither I took any classes on so I have no way of verifying if the formulas are indeed correct. The important part they were trying to prove mathematically is highlighted red concerning rod length which also determines important behaviors.

The other thing I need to know, is what part(s) of this are important to save?

I'm not out to copy their website word for word, I'm out to save the critical parts like formulas and behaviors that I highlighted in red.

Is the calculus formulas useful?

I suspect a chunk of this is wasted space and not really useful in the build/calculations department.

There is more left to salvage that goes with this. Will post it up tomorrow methinks. The dynamic compression page.

Getting the formatting of the formulas correct is a royal PITA and took a few tricks up my sleeve to do and a butt load of time.

 

Wow. This is great! Very clearly laid out.

Thanks for the props - but I really didn't do much! I will look into these formulas in greater detail soon as I get a chance to sit down at the computer for a while.

 

DYNAMIC COMPRESSION AND CAMSHAFT SELECTION

Dynamic compression is the effective compression ratio of an engine due to intake valve closing. Unless an intake valve closes immediately after BDC (bottom dead center, bottom of the stroke) compression is delayed, as the cylinder is not sealed. Exhaust intake opening does not effect compression as it always occurs after top dead center (ATDC). For this reason it is important to choose a cam that has an intake closing point ABDC that keeps the dynamic compression at a reasonable level.

Too high of a dynamic compression leads to detonation and loss of power. Too low of a dynamic compression robs power. In effect you want to run compression as high as possible without detonation.

With most modern import's this means you want a dynamic compression between 9.0:1 and 11.0:1 for a NA motor. For most domestic engines you want a dynamic compression between 7.5:1 and 9.5:1. The difference between import and domestic compression requirements has to due with head design. Most imports have more efficient head designs that allow higher compression without detonation.

Higher compressions are possible if a higher octant fuel is used (ie 110 octane race fuel). Octane rating measures a fuel's resistance to detonation, higher octane equated to greater resistance.

Figure 1: Piston, Rod, & Crank Diagram.

R = ½ stroke

S = Stroke

L = Rod Length

P = Position of piston at ABDC

P = R cos(q) + (L² - R² sin (q)²)^½

B = Bore

A = Atmospheric Pressure = 14.7 psi at sea level.

To determine dynamic compression we determine the amount of compression that occurs after the intake valve closes. To do this we must first calculate the effective stroke. Once the effective stroke is found we can use this to determine the effective cylinder volume, volume of the cylinder remaining for compression. This effective cylinder volume allows us to calculate dynamic compression.

Effective Stroke (Se)

The effective stroke is the actual stroke travel during compression (after the intake valves have closed). To determine this we find the piston position (P) at intake closing (ATDC intake cam data) and then add ½ the stroke and the rod length to determine the remaining stroke.

For more information on how P is found please review the RS Ratio tech article.

Se = R + L + P

(1) Se = R + L + R cos(q) + (L² - R² sin (q)²)^½

Effective Cylinder Volume

To find the effective cylinder volume we use the displacement formula, but replace stroke with effective stroke. Note to determine cylinder volume we do not multiple by cylinder count.

(2) Ve = B²* (Se)*0.7854

Effective Compression Ratio (CRe) / Dynamic Compression Ratio

To determine dynamic compression we us the normal compression ratio equation, but replace cylinder volume with effective cylinder volume.

(3) Cre=(Ve+H)/Ve

H is chamber volume. Chamber volume is the volume of the head per cylinder minus piston volume (above deck height). If H is not known the following equation can be used to determine H from static CR and Cylinder Volume.

H=V/(CR-1)

See basic engine calculations for more information.

Measured Compression/Cranking Pressure

Compression can easily be readily measured. Compression is the cranking pressure of the cylinder and is directly related to compression ratio. The equation below gives you the expected compression for an engine with a specific compression ratio (CR).

(4) CP = CRe¹·² (A) - A

(Thanks to Panic, Mopar Tech Papers, for equation 4)

CAMSHAFT SELECTION

It is important to take into consideration compression with camshaft selection (and vice versa). For an import engine you want the dynamic compression to be in the range of 9-11:1. An engine with 11:1 dynamic compression in a daily driver fueled with pump gas (premium) is living on the edge. An engine with 9:1 compression will be slightly anemic and an engine with 11:1 compression will be living on the edge. If the engine has been bored-out (reducing sleeve width) for added displacement I suggest running lower compression that with stock sleeves.

In the chart below is the dynamic compression of a number of Honda B series motors.

BLOCK	CAMS	Dynamic CR	Static CR
B16A	Stock	9.39	10.2
B18C1 (GSR)	Stock	9.24	10
B18C5 (ITR)	Stock	9.57	10.6
JDM ITR	Stock	10.03	11.1

As you can see the dynamic CR are in the suggested range of 9-11:1.

Some people suggest choosing a camshaft and piston combination that creates the highest cranking pressure and that ideally has a cranking pressure in a magic number range 15-20% higher than the stock. This methodology is flawed for several reasons.

1) Camshafts effect much more than compression and cranking pressure. Duration and lift of a cam greatly effect engine performance. A higher cranking pressure doesn't always mean greater performance.

2) Cranking pressure is a measure of compression ratio (see equation 4). There is nothing magic about 20% higher than stock cranking pressure. In general this MIGHT give you a dynamic CR near 11:1, but often this isn't the case. Choosing a CR higher than 11:1 based on cranking pressure increases is dangerous to you motor.

3) Increased CR (and cranking pressure) increases HP, but the gains are by no means dramatically different within the CR suggested range. Increasing CR gains HP in diminishing returns. Expect 3-4 hp/l for every CR point increase between 9-12 or until detonation occurs.

BLOCK	CAMS	Dynamic CR	Static CR	Cranking Pressure
JDM ITR	Stock	10.03	11.1	219.1182762
JDM ITR	B16A	10.25	11.1	225.2860365
JDM ITR	Skunk2 Stg 2	9.82	11.1	213.2560485
JDM ITR	Skunk2 Stg 2	11.48	13	260.2461232

To help illustrate the point lets look at the numbers in the table above. The B16A cams in a JDM ITR motor give higher cranking pressures than any other stock (11.10:1 CR) configuration. Using the cranking pressure methodology would suggest that the B16A cams would make more power and is false. Both the ITR and the Skunk2 stage 2 cams make significantly more power in this set up than a B16A cam; the Skunk2 Stage 2 cams make approximately 20whp more. Further to reach the 20% increase in cranking pressure the dynamic CR would have to be raised to 11:5:1. This CR is too high to safely run on pump gas.

The proper method of selecting a camshaft is complicated and involves many factors including cam lift, duration, compression, and engine breathing. Compression to some extent is the smallest piece. As long as dynamic CR is between 9 and 11:1 the cam is appropriate for the application. For a race motor aim for 10-11:1.

BLOCK	CAMS	Dynamic CR	Static CR	Notes
B20	B16A	9.99	10.8	Aftermarket Pistons
B20	ITR	9.77	10.8	Aftermarket Pistons
B20	Skunk2 Stg 2	9.57	10.8	Aftermarket Pistons
B20	Skunk2 Stg 2	11.06	12.5	Aftermarket Pistons
B20Z	B16A	9.2	9.96	Stock B20Z block
B20Z	Skunk2 Stg 2	8.84	9.96	Stock B20Z block
B20B	B16A	8.42	9.09	Stock B20B block
B20B	Skunk2 Stg 2	8.08	9.09	Stock B20B block

I've used a B20VTEC (CRVTEC) hybrid motor to explain proper cam selection based on compression in the table above.

The top three rows show a B20 block with aftermarket pistons that raise the compression ratio. As you can see the dynamic CRs are all with in the suggested range. Since this is the case I'd look at other aspects of each cam and the breathing characteristics of the engine itself to determine which camshaft I'd use in this application.

I've marked the fourth row in yellow because the Dynamic CR of this combination is borderline. 11.06:1 dynamic CR is very high and this can be a big problem in a block like the B20. The B20 block is a special case in the Honda Blocks as it has relatively thin sleeves, a large bore, and a short RS ratio. The thin sleeves in combination with high side loads due to the short RS ratio, and the large bore make detonation FATAL in this motor. To run 11:1 dynamic CR in this motor would require very good conservative tuning. I've heard of several B20 engines cracking their sleeves at these CR's. As such I'd suggest not running higher than 10-10.5:1 dynamic CR as a limit for these motors unless you are an expert tuner.

The next two rows are stock B20Z blocks (98-01 CRV). The 5th row is running B16A cams, the 6th row is running Skunk2 Stage 2 cams. The 5th row shows that the B16A cams should work fine in this motor. The 6th row (in yellow) shows that the Skunk2 Stage 2 cams will show disappointing gains in this motor and that compression should be raised if you wish to use these cams. As a side note the Skunk2 Stage cams are also border line on this motor due to possible valve to piston contact.

Finally the last two examples (red) help to demonstrate why some mild B20 Hybrid motors make 150whp and others make 180whp. The B20B blocks (96-97 CRV) have a lower compression than the B20Z blocks. Using them with stock pistons in a hybrid motor means your dynamic CR is terribly low and the motor will be anemic. If you are going to build a B20 Hybrid with stock pistons use the B20Z block.

Throttle Body

The size of throttle body effects engine response and peak hp. Here are some simple rules.

Large cross-sectional area increases throttle response.

Large cross-sectional area increased high rpm torque.

Small cross-sectional area decreased throttle response.

Small cross-sectional area increased low rpm torque.

Note: that optimal cross-sectional area is dependant on HP.

The following equation is a good rule of thumb for throttle body diameter (~ +/- 5%)

D (?) = sqrt[(WHP x 4)/(AP x CR)]

sqrt = square root

D = diameter of throttle body (single)

WHP = wheel horse power

AP = Atmospheric Pressure (14psi)

CR = Engine compression ratio

Cross Sectional Area = pi*Radius² or pi*(D/2)²

Note: Some engines have dual throttle bodies. The cross-section area of the dual TB should equal the area of the single

in the equation above.

Intake Manifolds

Choosing the right intake manifold is key to the type of characteristic a motor will have. Things to consider are runner

cross-section area (intake velocities), runner length (Helmholtz resonance), and plenum volume.

Some simple rules:

Short Intake Runners help high RPM torque.

Long Intake Runners help low RPM torque.

Large area intake runners help high RPM torque.

Small area intake runners help low RPM torque.

Plenum volume should be kept between ~40-50% of engine displacement. Small plenum volumes lower the rpm of peak torque.

Larger plenum volumes increased the rpm of peak torque. A plenum that is too small for the motor will starve the engine

of air, a plenum to large will bog it.

For detailed information on these effects check out the physics of intakes below.

Physics of Intake Systems

Cold Air:

Cold air is denser air. Engines need two things to make power, air and fuel. For every 7 degrees (F) of temperature

difference air density changes by ~ 1 %. Having access to cooler air can have a huge effect on hp.

Acoustic resonance:

Sounds are waves in the air, specifically longitudinal waves. These type of waves cause pressure waves (90 degrees

out of phase with wave itself) that are directly proportional in strength to the amplitude of the sound wave. It is

possible to utilize this effect to increase intake air pressures within the intake tube. In most instances the

amplitude of a sound wave is not that large. During resonance however the sound wave amplitude becomes much larger

(and so does the pressure wave following it). During this resonance it is possible to gain significant amounts of

torque and actually create positive manifold pressure (above 1 atm). This effect is sometimes called the AEM HUMP,

where power spikes ~10hp for a couple hundred RPM around 4500 RPM in their cold air intakes. The resonance frequency

of a pipe depends on the length and the diameter of the tube.

Fn= n (v/4L)

n = wave number, v = velocity of sound (~340m/s), L length of pipe

Note: v the speed of sound changes with temperature.

The diameter of a pipe has a small effect on the resonance frequency as well. The sound wave behavior remains consistent with that of a pipe for ~.6 times the diameter (D). So

Fn=n*v/4(L+0.6*D)

For a 1-meter intake pipe this translates into a resonance frequency of 85 hz or 5100 RPM. The intake pipe reaches resonance frequency when the driving force of the system (the engine) is also at that frequency. This is RPM/60; rotations per second.

Ram air effect:

Ram air is the process of using the relative airspeed of the outside air to ram itself into the motor. This effect is negligible at low speeds, but at high speeds it can be utilized to a small effect. Since air is non-compressible at speeds less Mach 1 the pressure created by velocity can be calculated using Bernoulli's equation for dynamic pressure.

Static Pressure + Dynamic Pressure = Total Pressure

Dynamic pressure = ½ (p) V2

P = air density

V = Velocity.

This translates into:

P (psi ? above atmospheric) = (.0000176) V2

V = Velocity (MPH)

Venturi effect:

The venturi effect makes uses the laws of fluid dynamics, specifically the ?equations of continuity?. Most people think

of pressure with reference to Boyle?s Laws for pressures in closed containers; P1V1=P2V2 (P= Pressure, V= Volume).

This equation is for static fluids and gases. When gases are moving the rules change a little. For gases in motion the

amount of flow is constant at the inlet and outlet. This means the area of the inlet/outlet determines flow velocity.

This makes since and is often utilized by someone with a garden hose. Placing your finger in front of the hose creates

a smaller stream of water that is moving much faster.

Equations of Continuity Q = V1A1=V2A2 (non compression gas fluid)

Q = Flow Volume

V = Flow Velocity

A = Area

This is where things get ?weird?. Bernoulli?s equation tells us that the total pressure must remain the same at the inlet

and outlet. However we also know that the equations of continuity tell us that the dynamic pressure increases with air velocity.

Ps1 + Pd1= Pt = Ps1 + Pd1

Ps1 = Static pressure

Pd1= Dynamic pressure

However we also know that the equations of continuity tell us that the dynamic pressure increases with air velocity.

Dynamic pressure = ½ (p) V2

P = air density (constant)

V = Velocity.

This means that static pressure must go up within the intake tube when flow velocities drop and cross sectional areas get larger

(and vise versa). This seems counter intuitive, but here lies the fun of fluid dynamics.

So what does this mean and how can I use this to advantage?

Two ways:

1 Increase velocities to tune for a specific RPM. The Spoon Venturi intake manifold gasket (rubber chicken) has a reduced

intake manifold gasket size to increase velocities into the motor. Peak torque occurs when intake flow rates are ~240-260

ft/s. With a smaller runner area, peak torque is lowered. This helps low end torque on smaller displacement motors at the

expense of high-end power.

2 Reduce velocities to create greater pressure to help push air into a new camber. This technique can be utilized directly in

front of the throttle body to help flow air into the intake manifold. By increasing the intake diameter in front of the TB

(ie: intake diameter is larger than the TB) the pressure is increased before the throttle body opening. This increased

pressure can help move air through the TB and into the intake manifold. Larry Wilmer (of Endyn) suggests a 25% difference

between intake and throttle body cross-sectional area directly before the throttle body.

Helmholtz resonance:

Applies to intake manifolds. This resonance effect on the intake filter system is negligible. Helmholtz resonance applies to IM and

is due to the intake valves opening and closing. This results in pressure build-ups and releases and air traveling back and forth

with into the intake manifold. The effect is comparable to a mass on a spring, oscillating up and down. Runner length (and area)

and plenum volume on an intake manifold are sometimes designed to effectively use this pulsating effect to increase efficiency in

certain rpm ranges. In the 1960?s Chrysler did a lot of research in the area of Helmholtz resonators. A quick little formula was

found to find the approximate area for this effect.

RPM ~ 88400 / L

It is found from the following.

Speed of Sound (V)

V=sqrt(Y * R * T)

Y= 1.4 (specific heat of air)

R= 286m^2/s^2/k (gas constant of air)

T=273.15 k (0 c)

V=340m/s=13400?/s (~15 c)

V=355m/s=14000?/s (~40 c)

D = Distance

L = Intake Length

V = Velocity

T = Time

F = frequency

N = Round trips

RPM = RPM of Peak Efficiency

D = Vt

T = D/V

F = 1/T

D = 2L * N

RPM = (V/D)*60

RPM ~ 84000 / L (5 return trips).

The shock wave or wave front from the valve events oscilates/bounces back and forth several times (5 times) between valve events.

Ideally at "resonance" the wave front hits the valve as it opens.

That is the theory behind the N = 84000 / L ~ (V/2L *5)* 60 (depending on V - seed of sound)

If your intake manifold runner length (to the valve) it is possible to hit multiple ?resonance points?. With a runner length of 16?

resonance will occur at N-5,4,3 with peak efficiencies at ~ 5250, 6550, 8700 RPM. A change to 14? will produce peak efficiencies

at ~ 6000, 7500, & 10,000 RPM.

N is the ~ engine RPM of peak torque

L (?) is runner length from the IM plenum to the valve.

Short Intake Runners help high RPM torque.

Long Intake Runners help low RPM torque.

Intake velocities:

It is generally accepted that peak torque occurs when flow velocities are ~ 240-260 ft/s into the motor. These velocities depend on

engine displacement, flow area, and RPM. The following equation can be used to determine peak torque due to intake runner area.

Peak Torque (RPM) = 88200 * Area / Cylinder Displacement.

This is all the rest of the math based info from ftlracing.com

By the looks of things, ftlracing used to house a lot more and this is what is left. And it's hard to say how long the website is going to be around now that it has downsized and seems to be a near dead site. So copying the info here may just preserve it.

Also this should allow peer scrutiny as there have been some human errors here and there that have been corrected and hopefully with peer help will continue to get cleaned up.

I am thinking I also need to trim off all the blubber and just have formulas and foundation information for a general "worksheet". In the meantime it's here and can be corrected if anyone see's anything out of place that needs a fixin'

 

After a bit of research, I am skeptical on the validity of the formulas from FTL racing.

The last really complex formula has been broken down incorrectly as can be verified here at wikipedia:

Piston motion equations - Wikipedia, the free encyclopedia

Using the wikipedia formula, I was able to convert the 28.5° ABDC to be a position of 125.38mm. Up from the piston bottom position of 121.75mm as for when the start of the effective stroke is at when IVC occurs.

I also found a much better evaluation set of formulas with full demonstrations showing good and bad and a serious amount of explanation on victorylibrary.

The thing I struggled with at victorylbrary was locating the pertinent documents but google turned up the one I needed very well.

The one I'm talking about is cam timing versus compression analysis.

Cam Timing vs. Compression Analysis

It goes on to explain why just the DCR alone is not a good representation of the motor and doesn't necessarily dictate octane requirements.

The formulas are based in inches and everything Honda is mm. So I am going to sit down and play with the numbers and see what the conversions of the formulas need to be to be for metric.

Victorylibrary has some solid information and I think is a much better set of tools for engine builders to be.

It has also given me the idea that using the domed piston not only reduces the cylinder volume but the upped compression may not necessarily make up for the reduced intake volume.

Further research also shows that I would have to mill the deck significantly to bring back the quench area which as the PMS pistons are 0.030" below the deck and even with the 0.025" 2 layer Y8 gasket, the 0.055" quench area is a little too much to be useful.

Also the domed pistons look to mess up the turbulence of the squish pads to the center. I'd probably be better off with the PM8 pistons from the JDM D15B vtec motor but those are rare. The dish would provide a very nice center vortex of turbulence and unlike the dome piston, would aid in scavenging.

I think doing the V/P index from Victorylibray may just demonstrate the pros and cons to each setup and see if the domed piston will actually impede performance even though it's upping the compression ratio significantly. Less volume in even though squeezed more may not work out so well.

I am determined to use the 3mm longer rods as I want to shot peen them and balance the engine to safely go 8500 RPM rev limiter. The regular 137mm rods are considered safe to 7200 stock so I figure with ARP rods bolts and shotpeening should allow the longer rods to go up to 8200-8500 rpm safely. Or so I'm hoping.

 

After taking some time off at wrapping my mind around all this shtuff I sat down tonight and started dinkering a little bit with the formulas on Victory Library I posted the link for in the post above.

Anyways, it was showing a safe peak speed of forged pistons to be 5000 feet/minute.

I'm converting the formulas to metric which in turn also changes the constant used in the formula.

Anyways, I found that forged pistons will safely peak at about 9000 rpm. Thought great.

Next formula is on piston acceleration. They were saying that 100,000 feet per second per second is safe limit but will cause flutter on 1/16" (1.5875mm) rings.

This formula showed that even with the longer rod, max rpm is about 7000 to stay within the acceleration "safe" limit.

I know even 15 years ago, Honda/Acura has motors running at 8500 rpm stock. I had thought the Integra Type R was one.

Can anyone provide the stroke and the rod length of any of the stock motors that run to 8500 rpm prior to modification?

Or if not the stroke and rod length of these motors, can anyone provide the motors that run to 8500 rpm stock?

I'm looking to compare specs to see if there is any wiggle room in the "safe" margins provided by Victory Library.

Thanks in advance.

 

It's not Honda, but check out the old (90's-early 00's) Lotus engines. They are actually Toyota 4-cylinders tuned by Lotus and spin to 9,000 rpm stock. Might be worth your while to look at the internal specs on those

 

FYI. The Integra type R has the same internal stroke/rod length (and thus r/s ratio) as the regular integra GSR. Same block as well. The ITR crank is heavier and the rods are a different casting.

Another interesting comparison is between the B16A and the B16B. Same stroke in both b16s but longer rods (and taller block) in the B16B. Red line in the b16b is 8900 rpms where as the regular b16a were limited to~ 8000 rpms.

Another good comparison is the Honda s2000 engines. Earlier ones were 2.0 liters and revved to 9k. The later ones were stroked to 2.2 liters and the redline lowered to 8k rpms.

Using what Honda developed and used in their production engines and the formulas you posted should help you design what you need using both factual and theoretical data.

Love the tech here even though I understand less than 1% of.it lol. Hopefully will spark others to shake their thoughts. .

 

What you posted about dynamic CR and static CR should also. Help debunk the myth about a particular engine creating X psi on a compression test...
Camshaft timing, valve lash, temperature, and even the compression tester itself are variables in what psi a given engine will produce. Kinda like comparing dyno numbers.

Thanks for bringing some tech and theory back to the all motor section. Even though 90% of readers here may not understand it or post up, the view count alone tells you it's a perking interest.

I wish I could contribute more to your discussion But I'm horrible at math/algebra/ trig/calculus. Basically anything involving mixing numbers and letters haha.

 

Same Engine as the Toyota Matrix GTS and Celica gts of that time right? 1.8 liter engines making 100hp/ liter ?
(10 years after Honda reached it lol)

 

Regarding piston speeds and acceleration. The 1st gen S2000 engine (2.0l redlining at 9k rpms) was reportedly to have one of the highest piston speeds in a production engine. Others that were in the same range were like F1 ad Ferrari engines if I remember correctly. .
The later "improved" S2000 lowered their red line and stroked it I am assuming to create a more reliable longer lasting engine (and more mid range TQ that so Many people complained of)

 

From what I've read and learned over the years (not just Honda forums) domestic and German engines as well. The best piston design is a flat top. No dome in the way to prevent less flow, less obstruction to prevent good mixing of A/F combinations and even combustion pressure over the entire piston.

So basically you want a nice flat piston, good quench (close as possible with out contact I think) unshrouded valves to fill the chamber as best as possible ( keeping in mind the biggest restrictions of air into and exiting the CC are the valves themselves) , as even as possible air fuel mixture and and much cylinder filling as possible.
Getting everything to work together at the desired rpm is what it's all about. 100% + VE

 

Initial research is showing the Lotus Elise for the USDM used the Toyota 2zz-ge motor. I'm hoping this is on the right track as it's the only reference for Toyota engine tuned by Lotus (to get an additional 10 HP out of it) that I am finding.

I come up with information showing the stroke to be 85mm and the same 1.62 Rod/Stroke ratio. Based off of the math it looks like the rod length is 1mm longer at 138mm over the Z6 1.37...

Found a post on a toyoat site that said the same thing, rod length center to center is 5.433" which is the 138mm.

On that same site it said something about rod angle already has the motor maxed for a street machine so can't be pushed much farther as can be seen in the quote below:

Then another snippet about the engine concerning square bore:

Looking at the "square" concept, I don't think that's the real deciding factor of RPM as much as a guidline of typical behavior of the motor. And as always there are exceptions.

So the toyota engine has 1mm longer rods, and .5mm larger stroke. And seems okay being taken to at least 7800 rpm and I think the lotus goes to 9000 you were saying.

Looking at the JDM B18C Type R, 8400rpm:
Rod Length is 137.9mm (just shy of the 138mm)
And a large stroke of 87.2mm
With a really low (stroker type engine) rod/stroke ratio of 1.58.
that must put some scary forces on the bottom end.

So I'm not sure how to digest this info... I suspect the rings on the pistons must be thicker to avoid flutter and I don't think the pistons are cast or the rods for that matter. I'm not even sure what other aspects need to be considered for those kinds of stresses at the high rpm range (8500).

 

What's also interesting is that the K20 is a perfectly square engine at 86mm x 86mm.
So after years of success with the B series engines and the aftermarket pushing the performance envelope, they decided square is the best way to go. Like what the B17 engine was 20 years ago... and what Nissan has been doing for years also. Hmmm

 

Thank you, thank you. Missed your posts when posting... Great bits of additional info.

I'm learning quickly that it is a very complex and intricate process.

Your additions are greatly welcomed and I'm thankful to have more to look into to get a better understanding.

 

It is. Very complex and intricate. That's what I love about it.
I've also came to learn that the formulas get you pretty dam close but there's always some unknown/ unexpected factor that throws it off a little.
A wise man once told me. "Sometimes it's better to be lucky, than it is to be good"
When it comes time for me to design an intake system for my build, you can expect me to be hitting you and this thread up for some guidance ha!