Rotary Car Club - Intake Design Exploration

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vex	01-25-2010 01:17 PM

Intake Design Exploration

So I've decided I want to produce a custom intake manifold, but I have no idea where to begin. Anyone have any links or words of wisdom on intake design? Plenum design, runner lengths (variable and fixed), pulse tuning, air intake velocity effect on torque and horsepower, etc, etc, etc. I'm looking for it all. I have a rough design in mind, but I would like to refine it and make it better and eventually using CFD to ensure the best possible outcome before producing it in real life.

So... Let 'em rip!

http://books.google.com/books?id=DoY...age&q=&f=false

need RX7

01-25-2010 03:50 PM

http://www.rx7club.com/showthread.php?t=199788

Quote:

Originally Posted by rotarygod (Post 1904435)

Well everyone always wants to know the mathematical way to figure out intake runner lengths so they can design their own manifolds. There are so many things to understand and the math seems to go on and on forever. Since there are books dealing with the subject in great depth I'll just get to the simplified math so you can figure out the perfect legth for your port style. Using this formula you will also learn why a halfbridge or full bridgeport engine utilizing the stock intake manifold has no top end but great midrange power. Here it is:

L= ( (1080-EPD) X 650 / (RPM X RV)

L= Legth of the runners. This is your answer.
EPD= Effective Port Duration (how long they are open for)
RPM= The spot where you want peak horsepower to be. (If you still have the stock gear ratio transmission and this is a streetcar there is absolutely no point in making this anything other than 6500-7500 rpm.)
RV= This stands for Reflective Value. The pressure wave reflects back and forth several times inside the pipe. For the intake the second wave is best so use the numeric value of 2. For a carburated car use 3 or 4 since the manifold may be too long. If you are figuring out exhaust length use 1. this will give you the proper length for a short primary collected system. If you want a long primary system, take the short length and multiply it times 4. OK lets plug in some numbers to prove this.

Let's look at just the primary ports in an '86-'88 n/a 6 port engine. The ports open at 32* ATDC (after top dead center) and close at 40* ABDC (after bottom dead center). We use 720* as our base point to start figuring out EPD. Since the port opens after TDC, we subtract 32* from 720* to get 688*. Since the port closes 40* after BDC, add 40* to 688* to get a total EPD of 728*. You now have one number to plug in to the above formula! So far the formula is (1080-728) X 650= 228800. Now we need to know what to divide this by. Since the '86-'88 n/a 6 port engine had a power peak of 6500 rpm this is what we will use for this example. Also since I said the second reflected wave is best to use, use the numver 2 for RV. There are the rest of the numbers for you. Take 6500 X 2 = 13000. Now we have 228800/13000. The answer; 17.6" There is one other thing to consider though. The reflection doesn't take place at the very end of the intake runner pipe but rather at a distance 1/2 the diameter of the pipe out away from the end of it. (what the hell did he just say?) Go back and read that slowly. Since the primary intake runner is 1 1/8" in diameter we must subtract half of this value from the length of the intake runner. .56". According to the calculations, the proper primary intake runner length for a stock port '86-'88 n/a 6 port engine is 17.04" The actual length as published by Mazda is 17.1"!!! Holy crap it works!!! The slight difference can be attributed to several small things. First we did not account for how fast the ports open and close. A peripheral port opens faster than a side port. accounting for this would give us shorter runner lengths. We also didn't account for the distance within the rotor housings that the air has to travel. This would add to our length. Basically these numbers almost cancel each other out so I don't worry about it. If you want to know how to figure out down to the last thousandth of an inch it will take some studying. If you look at the above number though we are within .06" of an inch from actual. Close enough. Your altitude where you live will affect it much worse than that.

If you need port timing specs for different year models I do have them as well as some of the Racing Beat port template specs and peripheral port specs. You will see something very neat when you type in the specs for a streetport but retain the factory intake manifold. Your horsepower peak will get lower! For my streetported GSL-SE I actually need a manifold with runners a little over 1" shorter just to retain the stock 6500 rpm peak number. This gets even worse with a half bridge or semi peripheral port. Many people think that since they ported it, their top end power falls off faster because lack of fuel. That may play a small role but they run rich up there anyways so big deal. The real problem is that the stock manifold they are using is too long. BDC if you are reading this it explains why Tony's half bridgeport car has much more midrange power than the streetport but falls off hard on the top end. Before you spend money on a new turbo, build a new intake manifold! Yours is too long. Yes it still matters on forced induction cars too!!!

If we wanted to tune the above engine to have peak power at 7000 instead of 6500 rpm the length would change to 15.75". For 8000 rpm it would be 13.75". Cool huh! Your low end will suffer the higher it is tuned and your power band will get narrower. You may quickly get out of you gear ratio range in which case all your new found power is worthless. If you used a Guru Racing transmission that has a shift point centered at 9500 rpm then you would want to tune for peak power at 9000 which would give a runner length of 12.15". The intake runner diameter may not flow enough air due to size though so this is something else to consider.

Now I suppose you want to figure out proper plenum size. I'll get into that later. I will just say that it gets SMALLER as the rpm limit rises! I'll let you all sleep on that one! If you have any questions about your porting style shoot me a pm and I'll help you out best I can. Have fun!

Cheers! :beerchug:

Fred

vex	01-25-2010 05:32 PM

That works well, but what about the other stuff? Plenum design and so forth? (The RCC is a different beast when it comes to math--we like the nitty gritty... I'm actually going to be looking into a book that deals specifically with this question I posted, though it's out of print)

I'm going to check out this book:
Theory of Engine Manifold Design: Wave Action Methods for IC Engines
ISBN: 0768006562

vex	01-28-2010 06:37 PM

Okay, so that book is okay... but doesn't cover anything specifically with rotaries. I saw maybe 6-7 books total in my schools library concerned about rotaries, and they were all from the 70's... I may have to start looking up more ASE articles for it... sigh, not what I wanted.

RotaryProphet

01-29-2010 09:06 AM

Rotaries aren't that special; treat them as a two stroke two cylinder of stated volume (ie, 1.1 or 1.3 literss), or as a four stroke engine with twice the RPM. Port timing info is available, and with all of that info, you should be able to calculate runner size and length and plenum volume, as well as throttle body size for your given peak torque/horsepower point.

As an interesting note, by using exhaust and intake runners a couple of inches longer on one rotor than the other, each rotor will have a different peak power point about 500 or so rpm apart, which leads to a wide peak power. Helpful if you have a particularly peaky motor like a P-port or a big bridge.

PercentSevenC

01-29-2010 02:40 PM

Here are some more relevant posts by rotarygod:
http://www.rx7club.com/showthread.php?t=94362

Quote:

Originally Posted by rotarygod (Post 844653)

I have a custom built aluminum manifold on my 2nd gen. It has a 100 cu in plenum volume. The intake runners are slightly larger than the lowers but there is a taper down to their size. This is done to broaden the torque curve since the air accelerates as the volume decreases. I also have a 75mm Mustang throttlebody on it with my blowoff valve machined about a quarter of an inch in front of the throttle plate. This engine was designed for big power but still be street drivable. I'll have to post some pictures later this week. The mathmatical formulas are hard to apply to a rotary. Keep in mind that on a piston engine the air is stagnant longer in the intake runners than they are in a rotary. On our engines the intake ports are only closed for a very short amount of time. On a piston engine, unless it is a 2 cycle, the air in the ports is not moving until every other time the piston is at the top. Because of this the effect of plenum volume changes in relation to piston vs. rotary. A bigger plenum volume will raise the torque peak. Port runner length is also dependent on port runner size. The effects are due to a volume/velocity relationship. Theoretically peak power (n/a) is achieved at whatever rpm the engine is at when air velocity entering the engine is at .6 mach (60% the speed of sound). This is however offset by runner length as a runner of too much length can kill power. Its almost too complicated for me to get into without writing a huge book (already several out there on this subject) or giving a night school class. I'm afraid something I write will sound contradictory to something else. Lets just say my setup was trial and error and it works damn well! Oh btw a pp engines intake is always open. When the apex seal crosses the port it is briefly open to 2 chambers. Therefore it never closes.

Fred

vex	01-29-2010 05:42 PM

Quote:

Originally Posted by RotaryProphet (Post 108365)

So Plenum is only concerned with volume and not geometry?

Quote:

As an interesting note, by using exhaust and intake runners a couple of inches longer on one rotor than the other, each rotor will have a different peak power point about 500 or so rpm apart, which leads to a wide peak power. Helpful if you have a particularly peaky motor like a P-port or a big bridge.

But would that put a moment on the E-Shaft and cause additional wear or are the forces congruent with doing that negligible?

vex	01-29-2010 06:46 PM

So I've started a preliminary design of the intake. And Suddenly everything I learned in Aero/Hydro dynamics is making sense to me and how I can mathematically apply what I learned. I may have a V2 of the manifold done by the end of the month. This should be interesting... :D

RotaryProphet

01-30-2010 08:45 AM

Quote:

Originally Posted by vex (Post 108414)

So Plenum is only concerned with volume and not geometry?

In as much as flow patterns are concerned, geometry is important, but as long as it's not shaped so as to -prevent- flow, you should be alright.

In an application where the throttle body is on the side of the plenum (facing forward in a rotary application), it's best to use a plenum four to six inches longer than the distance from the front of the first runner to the back of the last runner, and taper the plenum the entire way. This helps the air "slow down" earlier in the plenum instead of wanting to slam into the back wall, and helps the front runners breath. In a setup where this isn't done, the rear-most runner tends to ingest the most air, and the front-most one (from the throttle body's perspective) tends to ingest the least.

In a side-facing throttle body, the best solution would be a setup that tapered in both directions, with the widest portion in the middle, and some sort of diffuser to help the air with it's right-angle turn into the plenum, going either right or left. However, in a rotary specifically, since the middle ports ingest less air anyway, you can get very good results with a simple tube with a throttle body stuck on it, and let the "bad" shape direct the majority of the air into the big ports, where they're needed.

Also it's important to bell-mouth your runner entrance from the plenum, or better yet, use short velocity stacks actually sticking into the plenum.

Quote:

Originally Posted by vex (Post 108414)

But would that put a moment on the E-Shaft and cause additional wear or are the forces congruent with doing that negligible?

Considering that only one rotor is ever firing at a time anyway, you've always got more force (significantly) from one rotor than the other. The difference of maybe 10 horsepower at max between two rotor firings is negligible in comparison. No different than having a slightly low compression rotor and a good one.

It's really much more useful in V8 applications, where each pair of cylinders 360* off from each other is setup with peak power 500 RPM off from each other, creating, say, two cylinders making peak at 4500, two at 5000, two at 5500, and two at 6000, creating a very wide power curve. Obviously you're trading a reasonable amount of peak power (up to maybe 30hp) for this much wider band, but in situations where that's desirable (notably rally and drift racing, and some road racing), this is a good way to help. To really pull it off you need individual cylinder fuel and spark control, however.

vex	01-30-2010 11:46 AM

Alright, I think understand what you're saying concerning the Plenum design. Shouldn't be too hard to do a digital mark up in a few. As for the velocity stacks I was already considering doing that.

My question however is for turbo charged applications does wave tuning do that much to begin with? I'm curious because if I'm understanding it correctly the positive pressure comes on (for me anyways--Turbo 6PI) around 2k RPM. If i'm getting positive pressure that quickly, no matter what wave I tune for with the intake runner lengths I'm going to end up with more pressure than the wave could shove in by itself (even when it's not during a compression wave).

If my thinking is correct then I could be able to have rather long runners and be fine power wise. My concern from this however is will throttle response be adversely affected by having abnormally long runners? As it stands right now I'm thinking about keeping the runner length the same as a stock NA (roughly 17" or so), but directed much differently so the air needs to only take one continuous turn once inside the manifold.

I think the best bet for me (and everyone else who follows this thread) right now would be to focus on one stage at a time. For now lets focus on Plenum design and worry about intake runners later:

If I'm understanding you correctly you are telling me that a diverging-converging
(Air-> TB:<Plenum>) nozzle design will work best for when the throttle body is placed on the longitudinal axis of the Plenum. Lets focus on this setup as it seems the easiest to manufacture and produce in ones garage.

As air inters the divergence portion of the plenum the air will slow down according to thermodynamics:

A/A*=1/M[(2/(k+1))(1+(k-1)/2*m^2)]^((k+1)/(2*(k-1)))

Since it's air, k=1.4 and M<=0.6 the formula will give you A/A* for the divergence. (Note: A- Area when gas inters, A* When gas is at M speed)

If the pressure drop is significant enough the temperature will drop, but flow speed will suffer as it drops down in mach number. Knowing what the temperature will drop to we can solve for velocity using

Ve=sqrt(k*R*Te)

Note: There is a conversion factor in here (and this will end up in metric units)

Now you mentioned something about a baffle or is that not needed in a rotary application? Do we even need a Divergence-Convergence Plenum, or is it just a simple matter of getting a big enough pipe and sticking a throttle body on the end of it?

Aside: Can anyone scan a lower intake gasket for a 6PI and take a single measurement for me? I wish to be accurate and start a digital construction of the intake system so when I have access to CFD I'll be able to accurately see where potential flow issues are. Thanks

NoDOHC

01-31-2010 12:10 AM

Be careful applying Speed-of sound math here... If you do a good job on the plenum, you should be able to mostly ignore the effects of the compressibility of the air. Most of your resonance tuning is due to the Helmholtz effect (AKA: organ pipe, more of a dynamic systems model than anything to do the compressibility). The air in the plenum should not be moving anywhere near the speed of sound.

Don't think about this too hard man, intake manifold are simpler than they would initially seem. Don't ever try a dynamic model on a manifold unless you are a glutton for punishment (I have tried it, it is not easy).

Basically, your air velocity will follow the offset-sinusoidal waveform typical of an infinite-length-connecting-rod reciprocating engine (or a rotary, which has similar characteristics). The pressure drop at each transition is easily determined by using the lookup tables in the back of your fluids book, no difficult math required. Basically, you can get easy cross-sectional area requirements by taking the peak flow into the chamber and dividing it by the desired velocity (no rocket science there).

With the plenum, everyone has their own idea as to how bast power is obtained. I won't take to time required to explain my opinion on that. As I said before, you can easily find the flow through any given portion of the manifold at any given time with reasonable accuracy.

I will venture to say that I have seen tapered plenums, log plenums, cross rams, tunnel rams, inboard velocity stacks, tapered tubes, straight tubes, etc. in operation and I have not seen the simple log with beveled, constant cross-sectional area runners beat yet.

I hope this helps some. I know that math is awesome, but don't let it bog you down. Seriously, I found that going by my intuition and what feels right is often better than trying to crunch crazy numbers, there are too many x-factors to make any good simulations given the typical person's toolbox.

Edit: I hunted high and low for an intake manifold gasket and only succeeded in concluding that it is high time to clean out the garage (I know I have two brand-new ones, somewhere).

Can I take a scan of a LIM for you? (I can find that...) What measurement do you need?

vex	01-31-2010 12:23 PM

Quote:

Originally Posted by NoDOHC (Post 108547)

Understood. I doubt very much that the velocity of the intake stream will be higher than .6M which minimizes compressibility of the air. If I maximize intake velocity then wouldn't that mean that I have the ability to pull higher velocity air into the intake stroke for a higher torque curve? As for the Helmholtz effect, would actually tuning for that even though I'm turbo'd be beneficial? I suppose for cruise when I'm not in positive pressure it may be beneficial though I'm having a hard time reconciling the previous posts mathematics with my engineering brain (units don't add up). Is there a full formula somewhere that I may be able to tinker with?

Quote:

Don't think about this too hard man, intake manifold are simpler than they would initially seem. Don't ever try a dynamic model on a manifold unless you are a glutton for punishment (I have tried it, it is not easy).

I have access to CFD software and already have the wherewithal to create digital representation of the manifold in a few hours time. From my understanding once I have that all done and taken care of it shouldn't be more than a few more clicks and having it run to numerically solve the flow potential.

Quote:

Basically, your air velocity will follow the offset-sinusoidal waveform typical of an infinite-length-connecting-rod reciprocating engine (or a rotary, which has similar characteristics). The pressure drop at each transition is easily determined by using the lookup tables in the back of your fluids book, no difficult math required. Basically, you can get easy cross-sectional area requirements by taking the peak flow into the chamber and dividing it by the desired velocity (no rocket science there).

Don't have a fluids book :( I have an aerodynamics book, a couple ocean engineering books, and thermo book. I'm currently taking a compressible aero course but no fluid tables for different offsets.

Quote:

With the plenum, everyone has their own idea as to how bast power is obtained. I won't take to time required to explain my opinion on that. As I said before, you can easily find the flow through any given portion of the manifold at any given time with reasonable accuracy.

I will venture to say that I have seen tapered plenums, log plenums, cross rams, tunnel rams, inboard velocity stacks, tapered tubes, straight tubes, etc. in operation and I have not seen the simple log with beveled, constant cross-sectional area runners beat yet.

I imagine that the velocity increase from velocity stacks, tapered tubes, etc wouldn't yield a high enough velocity increase to matter.

Quote:

I hope this helps some. I know that math is awesome, but don't let it bog you down. Seriously, I found that going by my intuition and what feels right is often better than trying to crunch crazy numbers, there are too many x-factors to make any good simulations given the typical person's toolbox.

I'd hate to be argumenitive (and I don't want to come off like that) but I do not have the typical person's toolbox: CFD, Numerical analysis, and a hand full of individuals that have 800+hp cars. I've bounced my thoughts and what I am thinking off of them and they seem to be of the mindset to use the CFD software to ensure proper air distribution to all runners is key in plenum design. I appreciate the information that you're giving to me and by no means am discounting it off hand, i'm just trying to learn as much as I can in as little time as possible.

Quote:

Edit: I hunted high and low for an intake manifold gasket and only succeeded in concluding that it is high time to clean out the garage (I know I have two brand-new ones, somewhere).

Can I take a scan of a LIM for you? (I can find that...) What measurement do you need?

It's okay, I don't want you to break you scanner. I may actually be getting a LIM in the not too distant future to take the measurements off of myself. But if you're wanting to do it you'd just need to scan the profile of the mating flange, and measure say a mounting hole diameter. This sets a scale for a digital copy and will allow me to pull the measurements off the copy with out worrying if they're right.

Basically Dreal/Dscale=Dreal/Dscale: 10.5mm/1.25mm=Dreal/8.9mm; Dreal=74.76

NoDOHC

02-01-2010 11:05 PM

1 Attachment(s)

CFD software is outside my experience, but if it is capable of simulating the flow in the manifold accurately (assuming that it is given good data) you have an excellent opportunity to maximize your learning as you go, while using the resources that you have to their fullest potential.

As you said, you do not have the typical toolbox.

Helmholtz tuning works for any level of boost (it is basically the natural frequency of resonance of the fluid system). The speed of sound varies with density of the charge, so you will have to that into account on the Helmholtz equation.

You can write the Helmholtz equation most easily in terms of the resonance frequency (First equation) Solved for L gives the second equation. This expects a uniform cross-sectional area for the runner, as any changes in velocity will create additional and possibly conflicting pressure waves.

(a = speed of sound, V = velocity in the runner at time of wave excitation, L = Runner length from source to plenum, A = cross sectional area of runner, f = frequency of resonance (which is related to engine speed, obviously).
I hope this helps some.

With the tools at your disposal, this should be one awesome manifold.

The offset sinusoid expresses the chamber volume as a function of E-shaft rotation (the period is 270 degrees, the amplitude is 20 in3 (327 cc) (654 cc peak to peak). Taking the derivative with respect to time requires that the x axis be in time units (pick an engine rpm). This will give you the rate of change of chamber volume with respect to time, which should give you a good velocity characteristic (given port cross-sectional area). Hopefully this will be good enough input data to get a reasonable approximation of how the manifold will flow.

Keep up the good work!

vex	02-01-2010 11:10 PM

Quote:

Originally Posted by NoDOHC (Post 108744)

:o16::drool5::o13: Thanks! Those look much, much, better than what I was looking at a few minutes ago!

vex	02-02-2010 01:58 PM

Okay, I'm working with the formula's you gave me and I've got it narrowed down a little bit. I have found at least two parts where the pressure will affect the runner length:
The amount of air displaced (when calculating the velocity in the runners), and the speed of sound propagation (sqrt(gamma*R*T)). Temperature and Gamma may change with an increase in pressure. I'll post up the full formula I have when I've got the little details worked out.

vex	02-02-2010 04:32 PM

1 Attachment(s)

so... my original understanding of the formula was flawed so I did a little research, properly solved for L and then plotted it as a function of RPM. Here's a picture of the plot I was able to derive.
http://www.rotarycarclub.com/rotary_...6&d=1265146172

I also had the program pump out the minimum length at an 8k redline. Here's what I had the program pump out:

Quote:

>> runnerlength()
Assumed redline of 8000
Min Runner Length: 0.162229 m

Consequently I think it will require a little tweeking in order to be extremely accurate. But assuming that there's about 10 cm or so inside the engine itself it makes absolute sense that a 11 cm intake manifold at redline is so small... but who knows.

So what do you guys think?

(also this is for a runner diameter of about 4 cm)

tervo rodriguez

02-02-2010 09:38 PM

here is a look at my custom intake that positive pressure comes in at 2200k w/webbers dcoe 45 on 13psi boost at 3200k and puts out 322 to the wheels 360hp on my raceported 12A http://www.rx7club.com/picture.php?a...ictureid=13149 http://www.rx7club.com/picture.php?a...ictureid=13146 http://www.rotarycarclub.com/rotary_...pictureid=1571 http://www.rx7club.com/picture.php?a...ictureid=13145

Quote:

Originally Posted by vex (Post 108493)

A/A*=1/M[(2/(k+1))(1+(k-1)/2*m^2)]^((k+1)/(2*(k-1)))

Ve=sqrt(k*R*Te)

vex	02-03-2010 12:21 PM

1 Attachment(s)

Quote:

Originally Posted by tervo rodriguez (Post 108817)

here is a look at my custom intake that positive pressure comes in at 2200k w/webbers dcoe 45 on 13psi boost at 3200k and puts out 322 to the wheels 360hp on my raceported 12A

Your plenum and intake are simple enough. Though I don't think your runners can readily be seen in the pictures you provided (since you're running a carb the runners would be underneath the carb).

Though I must say, I really like the way your plenum is set up. It looks somewhat similar to my first attempt to digitally model the intake, though I didn't place the blow off valve on the plenum. I like the thought of having the vac/boost lines coming off the plenum. Do you run into any issue with pressure sensing with a turbulent flow? I suppose you really wouldn't because stock does the same thing.

As for the code generation for calculating intake runner length and diameter I have some good news to report:

http://www.rotarycarclub.com/rotary_...7&d=1265217595
I've also included the code below for further investigation for users who wish to run it themselves and maybe even critique my math. I wrote it in matlab which allows easy 3d surface rendering accurately.

Code:

function []=runnerlength()

gamma=1.4;

R=287;%J/kgK

T=293.15;%K



a=sqrt(gamma*R*T);%M/S



redline=8000;%rpm redline

v=0.0013;%m^3

r=linspace(0.001,0.015,100);%m

A=@(x)pi().*x.^2;%m^2



rpm=linspace(3000,10000,100);

f=@(x,y)a.^2.*A(y)./(4.*(x./60).^2.*pi().^2.*v);

z=zeros(length(rpm),length(r));

for i=1:1:length(rpm)

    for j=1:1:length(r)

        z(i,j)=f(rpm(i),r(j));

    end

end



surfc(r,rpm,z)

colorbar

xlabel('Radius, r (m)')

ylabel('RPM')

zlabel('Length (m)')

vex	02-04-2010 11:35 AM

I'm surprised... No critiques or suggestions on improvement? Is my math right? Do the results look accurate?

What runner diameter should I consider accurate? I've been leaning towards 3 to 4 cm in diameter, but if I wish to tune for 3500-5000RPM in peak torque I may run into problems actually fitting the amount of pipe for the runners.

I'll be fooling around with the code (different temperatures, narrowing down the diameter to use, etc) and see if I can't get a finalized result for everyone's benefit. This should be quite helpful for individuals (and vendors) who wish to design their own intake systems.

Consequently here's my next round of questions with regard to runners:
What are the benefits/drawbacks of having multiple diameter runners--Having primaries one diameter and secondaries another? What about tapering? Venturi/Velocity Stacks and their affect on pressure drops, and other flow conditions?

This should prove fairly interesting and thanks everyone for their help and support in this thread thus far! Hopefully others are learning as much as I am!

scotty305

02-07-2010 04:17 AM

Nice Matlab code, I used Matlab in school but never got around to using it for fun car stuff like this.

I have one comment and one question:
1. I knew some people who built a variable-volume intake plenum for dyno testing... it was essentially a box with fixed runners and a sliding lid (the throttle attached to the lid), sealing it was difficult but slightly easier than fabricating multiple plenums.

2. Is there a way of predicting the relative power gain or loss due to resonance tuning? For instance, if you could somehow predict that there would be a X% gain in power at the resonant RPM but a Y% loss due to some sort of antiresonance elsewhere, you might try runner lengths/diameters that not only maximize gains in your desired powerband but also minimize losses.

scotty305

02-07-2010 04:20 AM

Additionally, you might want to think about how you're going to test your setup. It seems it would be wise to use realistic ramp rates on the dyno (if the dyno software allows for this).

Quote:

I fully appreciate the benefits of step testing with sweep testing at the actual rate of acceleration of the vehicle. This can show completely different issues, and allow the separate tuning of load/rpm tables from acceleration enrichment. It can also help to separate engine loads from drivetrain rotating mass loads - both of which alter tuning requirements (both fuel/spark for engine management and pulse tuning in intake and exhaust systems).

An example of different results - Using step testing, I can develop a plenum where a specific volume shows a very nice improvement in output at higher rpm over a 1000 rpm window of the power band - without hurting low rpm output. However a sweep test will show a tendency to hurt lower rpm response and not influence upper rpm output nearly as much as the step test. On the track, this plenum causes the engine to be lazy with poor transient response. And a plenum 20%-30% smaller is actually considerably quicker and faster on the track - dyno results be d@mned.

The same things hold true in exhaust design.

All that matters is the track or street results - RIGHT?
Dyno results (HP/TQ numbers) are irrelevant other than for impressing uninformed customers, internet bragging rights, and the Engine Masters Challenge. They don't win races.

IMHO, both step and sweep testing need to be performed together. And all atmospheric conditions need to be taken, and corrected for, at the point of air entry into the vehicle.

Quote:

this is about actual results using several types of testing methods on both engine and chassis dynos, and comparing the results to actual performance on road or track, not theory.

On a quick accelerating vehicle, intake and exhaust pulse tuning does not have the same time to stabilize as it would with a step test spending 6-8 seconds at each rpm point. The difference can be very large. This is a well understood issue. Perform your own testing and see the results for yourself.

There is a big difference between tuning being "fine" and being right. To engine builders, racers, and customers trying to get the best results for their hard earned Dollar, this matters.

info from http://www.efi101.com/forum/viewtopic.php?t=4105

vex	02-07-2010 11:28 AM

Quote:

Originally Posted by scotty305 (Post 109162)

I didn't even think about doing a variable volume plenum. Honestly the requirements of such things for sealing it would be a little too much for positive pressure tuning. Consequently I have not seen any mathematical model to which I could digitally model the results for it.

Quote:

2. Is there a way of predicting the relative power gain or loss due to resonance tuning? For instance, if you could somehow predict that there would be a X% gain in power at the resonant RPM but a Y% loss due to some sort of antiresonance elsewhere, you might try runner lengths/diameters that not only maximize gains in your desired powerband but also minimize losses.

I doubt very much you could hope to tell how much torque you'd gain at a specific RPM. Based off the 26B the torque increase is a rather small percentage of incorrect length intake. There's also too many variables to stat that intake runners and a plenum volume will have this effect on a car: The difference between one car and another in power really comes down to the whole shebang. Intake, Combustion quality, Exhaust, rotating mass of the drive train, and even atmospheric conditions. You'd honestly have to hold 4 of those things constant just to test the effect of one--and even then I'm not sure that holding them constant would give realistic results.

scotty305

02-07-2010 03:26 PM

Sorry if that wasn't clear... the plenum volume wasn't adjustable on-the-fly, the 'lid' position was changed between dyno tests and this involved clamps, bolts and probably some sort of sealant. In hindsight I should have taken a closer look at the setup.

NoDOHC

02-07-2010 09:37 PM

Good call on the math error! I should have just given the original equation. Apparently the algebra is a bit rusty (I didn't think it looked right, but I couldn't find the formula solved for L in the book).
The natural frequency formula is correct though, I just checked it.

Anyway, I like where you are going with this. Something to remember on the Plenum is that the two rotors are 180 degrees out of phase and have a 270 degree intake duration. This means that both rotors take in air at the same time.

This is looking good!

There are more advanced formulas for varying cross-sectional-area runners, but I don't have the ambition to enter them into paintbrush and you can probably find them online anyway without any errors.

If you want to do the math yourself, it is very simple to draw it up as a dynamic system and then find characteristic equations for it. I say this because you are most likely taking a course in dynamic systems right now or in the near future.

The Plenum has compliance as it acts as an accumulator. The runner has a resistance (dissipative) and Reluctance (inertial) element to it. The port closing is the disturbance function. The system is a lot harder with changing area.

Anyway, you will find that the wave intensity is a logarithmic decaying function and that the time constant is related to the runner smoothness (resistance) and the air velocity (inertia). I can't find the equation right now, maybe it is in a book at work.

I will try to find it and let you do the Algebra.

vex	02-08-2010 11:14 AM

Quote:

Originally Posted by NoDOHC (Post 109231)

It happens to the best of us. I can't remember the times I've taken a test and screwed up some simple algebra and ended up with a horrible, horrible, horrible wrong answer.

Quote:

Anyway, I like where you are going with this. Something to remember on the Plenum is that the two rotors are 180 degrees out of phase and have a 270 degree intake duration. This means that both rotors take in air at the same time.

That's actually really helpful, but I think we need to adjust the angle of intake and have it based off of port size, and what not as you're going to get a larger intake with a bridge port compared to a street port

Well my class is finished, time to get back to what I was doing.

Quote:

This is looking good!

I like it a lot and I figure this will be an easy tool for other rotary (and heck, even them piston) guys to effectively make and design a proper manifold without the downfall of one too short or too long. If anyone is able I'd like to see someone make this into a web app that RCC can host on their page (I'll make a thread about it in lounge for those interested).

Quote:

There are more advanced formulas for varying cross-sectional-area runners, but I don't have the ambition to enter them into paintbrush and you can probably find them online anyway without any errors.

I don't even know where to begin with that. Could you point me in the right direction?

Quote:

If you want to do the math yourself, it is very simple to draw it up as a dynamic system and then find characteristic equations for it. I say this because you are most likely taking a course in dynamic systems right now or in the near future.

Actually I don't think I have any dynamic systems setup for my classes now or in the near future. This is why I'm doing a lot of it on my own. I can probably do something similar to what you're telling me in the schools CFD program (if I can ever get it running)

Quote:

The Plenum has compliance as it acts as an accumulator. The runner has a resistance (dissipative) and Reluctance (inertial) element to it. The port closing is the disturbance function. The system is a lot harder with changing area.

You're going to have to go into more detail if you're going to make me want to understand what you're saying. :lol: I think I understand, but I'm not too sure.

Quote:

Anyway, you will find that the wave intensity is a logarithmic decaying function and that the time constant is related to the runner smoothness (resistance) and the air velocity (inertia). I can't find the equation right now, maybe it is in a book at work.

If you ever come across it in the near future post it up (don't forget references).

Quote:

I will try to find it and let you do the Algebra.

DEAL!

vex	02-16-2010 08:36 PM

Crap, I think I made have figured out a mistake. Should the volume displaced be 1.3L or should it be half that?

vex	02-17-2010 03:05 PM

Okay, just double checked. The original formula I posted is correct.

We're not looking at the .0013/2 m^3 volume for displacement volume since intake is a continuous displacement of 1.3L (both rotors are 180* out of phase)

vex	09-29-2010 09:02 AM

http://www.rwdi.com/cms/publications/26/t09.pdf

^Some helpful rules of thumb for diffuser designs.

NoDOHC

09-29-2010 10:43 PM

I think I need to have a look in the attic, I can't find that book anywhere. From your other threads, it looks like you have done a lot more research than I can remember from old textbooks.

The intake is looking good!

I was thinking that the rotary intake strokes come twice as often as the piston engine intake strokes, meaning that the air had half the time to travel. The intake runner length equation may be off. I will try to get a chance to draw the dynamic model and solve the characteristic equations to see if I can derive the resonance frequency formula properly for a piston engine (it helps to stretch the algebra muscles every so often anyway, as they obviously atrophy). If I can, I will draw the model for the rotary and compare.

vex	09-30-2010 08:16 AM

Quote:

Originally Posted by NoDOHC (Post 129288)

Please do. I have since had that class and now understand what you're talking about. I honestly can't model a dynamic system like this without just retarding it down to a simple mass-spring-damper problem. I do believe the equation I have is correct, though I am of the same concern as you.

The way I looked at it is degrees of rotation compared one to the other. For 720 degrees of rotation of the crank you get 1 intake stroke (Just looking at a single piston), similarly for the rotary you get 1 intake stroke for 720 degrees of rotation (looking only at a single rotor).

I had a simple program at one point that showed the correlation between the intake stroke of the rotary and piston with respect to engine RPM. I'll see if I can dig it up and validate my mental experiment.

NoDOHC

10-07-2010 11:54 PM

Each Rotor has one intake stroke per 360 degrees of rotation.

The rotor completes 1 full revolution every 1080 degrees of eccentric shaft rotation. This is 3 intake strokes.

I think this will change the dynamics.

vex	10-08-2010 04:33 AM

Him... I don't know then. I may have to re-evaluate the length formula, though from my understanding and checking it against the stock length on the NA version of the FC it seems to be accurate.

NoDOHC

10-08-2010 10:15 PM

The nice weather will soon be over. This sounds like a winter project with a notepad and a calculator. I think I loaned engine dynamic analysis book to a friend, as I can't find it anywhere. If I can find it, I will try to derive the equations for a rotary engine (from the piston engine equation derivation explanations in the book). It has been too long since I studied any of this for me to trust my memory.

I know that the rotary engine has a 298 degree intake port duration each revolution, making the airflow into the two rotors overlap (one is still filling when the other starts). This is also why the engine is so smooth, as the power strokes of the two rotors overlap. The actual maximum to minimum volume cycle of the engine occurs in 270 degrees of eccentric shaft rotation.

Do you have an engine apart so that you can get a feel for what it does as it turns?

I can try to get pictures, but a photographer I am not (as I think is evidenced by many of my pictures on this site).

For what it's worth I think this development is very helpful to the community and I appreciate your enthusiasm and effort on the project so far.

Thanks!

vex	10-09-2010 02:09 PM

Quote:

Originally Posted by NoDOHC (Post 129919)

Yeah, sounds good... though with the colder weather comes more homework assignments for school for me.

Quote:

I know that the rotary engine has a 298 degree intake port duration each revolution, making the airflow into the two rotors overlap (one is still filling when the other starts). This is also why the engine is so smooth, as the power strokes of the two rotors overlap. The actual maximum to minimum volume cycle of the engine occurs in 270 degrees of eccentric shaft rotation.

That I understand, but for sizing the actual intake runner length you need only concern ourselves with one intake volume as the pulse from the closing of the intake port results in the wave propagation to achieve higher VE of the engine (in other words we don't really need to worry about the intake overlap charge). The wave travels the length of the runner at a specific RPM and results in a compression wave of air being sucked into the next rotation of the rotor.

Quote:

Do you have an engine apart so that you can get a feel for what it does as it turns?

No... but I do have a nice program to validate different cycles of various engines to ensure synonymous application of the formulas considered.

Quote:

I can try to get pictures, but a photographer I am not (as I think is evidenced by many of my pictures on this site).

There's really no need unless you really desire to.

Quote:

For what it's worth I think this development is very helpful to the community and I appreciate your enthusiasm and effort on the project so far.

Thanks!

As always I appreciate your input and help with my engineering cog turning exercises. Hopefully the final iteration of the formula will facilitate the proper lengths and diameters of the intake runners.

In the meantime however I need to do my boundary layer theory report.:o10:

vex	10-09-2010 08:54 PM

Of interest and the reason for using 0.0013 cubic meters:

Quote:

Originally Posted by wikipedia

There are various methods of calculating the engine displacement of a Wankel. The Japanese regulations for calculating displacements for engine ratings use the volume displacement of one rotor face only, and the auto industry commonly accepts this method as the standard for calculating the displacement of a rotary. However, when compared on the basis of specific output, the convention results in large imbalances in favor of the Wankel motor.
For comparison purposes between a Wankel Rotary engine and a piston engine, displacement and corresponding power output can more accurately be compared on the basis of displacement per revolution of the eccentric shaft. A calculation of this form dictates that a two rotor Wankel displacing 654 cc per face will have a displacement of 1.3 liters per every rotation of the eccentric shaft (only two total faces, one face per rotor going through a full power stroke) and 2.6 liters after two revolutions (four total faces, two faces per rotor going through a full power stroke). The results are directly comparable to a 2.6-liter piston engine with an even number of cylinders in a conventional firing order, which will likewise displace 1.3 liters through its power stroke after one revolution of the crankshaft, and 2.6 liters through its power strokes after two revolutions of the crankshaft. A Wankel Rotary engine is still a 4-stroke engine and pumping losses from non-power strokes still apply, but the absence of throttling valves and a 50% longer stroke duration result in a significantly lower pumping loss compared against a four-stroke reciprocating piston engine. Measuring a Wankel rotary engine in this way more accurately explains its specific output, as the volume of its air fuel mixture put through a complete power stroke per revolution is directly responsible for torque and thus power produced.

If I'm reading this correctly a proper displacement of 0.0013 cubic meters is achieved with 1 revolution of the eccentric shaft. Validating the Helmholtz equation used.

vex	01-25-2011 11:54 AM

So... I figured I should update this thread.

I'm currently going through design iterations of a new intake. As it stands I have balanced the runners with plenum geometry.

Some pointers to keep in mind when doing this yourself:

Plenum Geometry is very important with flow distribution to the runners
Hemholtz equation is useful for understanding the ideal length for a specific RPM
Plenum volume is important under natural aspiration criteria and less so in forced induction.
Slowing the intake charge once into the plenum will result in more uniform flow distribution of the runners.

NoDOHC

01-25-2011 08:36 PM

I should clarify a previous post. I re-read it and got confused myself.

Quote:

Originally Posted by NoDOHC

It is true that Rotor 1 (front) is taking in air during the same time that rotor 2 (rear) is taking in air (port active 288/360) = 80% of the time - therefore 60% of the time, both rotors are on an intake stroke. Unfortunately, I improperly computed the stock port open time for a 4-port (It is actually 288 degrees - 32 degrees ATDC and 50 degrees ABDC = 270-32+50 = 288). The 6-port is 268 degrees with ports closed (270-32+30) and 318 degrees with ports open (270-32+80).

My apologies on the misinformation.

The cool thing about the rotary engine is that although the port open time is overlap-equivalent to 2/3 the timing for a piston engine - 192 degrees for a four port (talk about a tame cam), the air has the same amount of time to enter the rotor as it would with 288 degrees of intake duration on a piston engine. Bear in mind too that well-ported ports on a rotary flow much better than an equivalent-sized valve (over 3 square inches of area between the ports would require one 2+" valve or 2 1.4+" valves). This is why the intake manifold is your biggest friend (or enemy).

Thanks for all your research in this vex!

vex	02-15-2011 03:12 PM

Time for another update (this will be used in conjunction with the NA to Turbo sticky found in the 2nd Gen forum):

To further understand the flow distribution to each runner in a manifold one needs to consider the geometry one has chosen. Flow distributions are going to be significantly different if one chooses runners 'inline' with the throttle body as compared to a 'non-inline' manifold.

A picture is worth a thousand words:
for a 'non-inline' type of manifold think of this:
http://www.roadraceengineering.com/p...e_manifold.jpg

for a 'inline' type of manifold think of something similar to this:
http://i15.photobucket.com/albums/a378/2a_ron/bay2.jpg

Both have their pro's and con's. For most people an 'inline' style manifold is going to be just fine for what they want. However there are some things to be made aware of if you're going to be making your own.

First and foremost, any changes to the stock manifold should be qualified. That is; use a CFD program to calculate how your flow is going to be affected. I'm not going to get into the nitty gritty of CFD programs or even how to run them, but suffice it to say that it's better to use one than to not (you can look up OpenFOAM for a free CFD program with a large support community--I however do not use it, so don't ask me). Once you have qualified your base design to ensure uniform airflow distribution between both the front and rear rotors look at the size of the throttle body. If you can't get the flow just right, take a close look at your intake velocity .

When using a CFD program, you may notice that some of the issues you encounter come from too fast of an airflow into the manifold (if you're doing boosted applications you'll need to calculate the volumetric flow rate coming through the throttle body). There are a few ways to overcome too fast of an airflow into the manifold. First and foremost increasing the size of the throttle body will inherently slow down the flow for a better and more uniform distribution. However, just upping the throttle body may not yield desired results. One must further look at how the air is expanding into the plenum. Without getting into a lot of engineering speak and mathematics; a gradual enlarging of the effective area of the flow will slow down the flow. Such an object is called a diffuser (no we're not talking about the thing that goes under the rear bumper--though they do the same feat).

Diffuser geometry is very tricky to get right. If, for instance, you enlarge the diffuser too quickly you will cause the air to separate (engineering speak: The boundary layer separates) from the wall of the diffuser causing turbulence and circulation (something we're trying to avoid at this point of the flow). If you carry on the diffuser for too long the boundary layer will begin to enlarge and separate; eventually causing circulation. This is, of course, based on the intake velocity. Slower traveling intake charge will be easier to expand than a faster moving charge.

For understanding the above here are some images of 75mm 'non-inline' manifold compared to a 90mm 'non-inline' manifold.

(reserved)

blaen99

06-09-2011 08:32 PM

1 Attachment(s)

Here's a very early beta of Vex's proposed program written in C# for ease of development.

It only accepts Vex's proposed metric values for now, but I need to verify if the values and math work correctly or not to go forward with further development.

vex	06-09-2011 08:55 PM

it looks like it's outputting the same values. Here's some tests for verification:

inputs:
r=0.019
rpm=6500
vol=0.0013
temp=293.15

Your output: 0.22178 m
My output: 0.2218 m (Due to format truncation).

r=0.019
rpm=3200
vol=0.0013
temp=293.15

yours: 0.91507
mine: 0.9151

Looks like your math is fine to me. Is there a specific error you're concerned with?

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