[GP125] All that you wanted to know on Aprilia RSA 125, and more, by Mr Jan Thiel and Mr Frits Overmars (PART 1) (Locked)

Sorry my mistake.

Can't a dyno bench measure negative values? Just pull the clutch after max rpm and then let the dyno turn the gearbox and measure the negative torque required. Of course this will get you the gearbox loss and not also the primary drive. So in the end we would still be guessing.

In that case the 17bar sounds correct.

Interesting if the highest NA 4 stroke BMEP in fact is the low reving Formula 3 engine. I can't think of any other with a higher value.

Jan Thiel a écrit:

I never talk about crankshaft HP as I have no means to MEASURE them!

I fully agree, Jan. I also prefer to talk only about measured values. But when calculating the mean effective pressure in the cylinder, by definition you must base the calculation on crankshaft power.

Mic a écrit:

Can't a dyno bench measure negative values? Just pull the clutch after max rpm and then let the dyno turn the gearbox and measure the negative torque required. Of course this will get you the gearbox loss and not also the primary drive. So in the end we would still be guessing.

Worse than that, Mic. The gearbox losses mainly consist of friction between the gears, which is a percentage of the transmitted power. Once you remove the engine power, this transmitted power is no longer present and the friction loss will be much smaller. And you simply cannot measure a power, neither positive nor negative, that is not present.

Frits Overmars a écrit:

With leaded fuel the compression ratio was 19,5; nowadays it is about 15.

Frits that is geometric compression ratio not trapped compression ratio, right?

Wow, only 65 bar peak cylinder pressure? I was expecting you to say something like 120 or 130 bar, especially with 2 bar trapping pressure and 17 bar BMEP. I thought I remembered Paul McEntee telling me he measured typically 100–110 bar on a GP engine at a time when BMEP was only 12 bar and trapping pressure was more like 1.6 to 1.8 bar.

Yes, that is geometric compression ratio, Brian. I find trapped compression ratio an unworkable concept. Moreover if you know an engine's trapped ratio, you still know nothing; you also need information about the exhaust port and even about the behaviour of an exhaust power valve.

I found this article about BMEP


Brake Mean Effective Pressure (BMEP) is another very effective yardstick for comparing the performance of an engine of a given type to another of the same type, and for evaluating the reasonableness of performance claims or requirements.

The definition of BMEP is: the average (mean) pressure which, if imposed on the pistons uniformly from the top to the bottom of each power stroke, would produce the measured (brake) power output.

Note that BMEP is purely theoretical and has nothing to do with actual cylinder pressures. It is simply a tool to evaluate the efficiency of a given engine at producing torque from a given displacement..

If you work through the derivation arithmetic (presented at the bottom of this page), you find that BMEP is simply a multiple of the torque per cubic inch of displacement. In fact, many talented people in the engine design and developmeny business currently use torque-per-cubic inch instead of BMEP, thereby avoiding that tedious multiplication.

A torque output of 1.0 lb-ft per cubic inch of displacement in a 4-stroke engine equals a BMEP of 150.8 psi. In a 2-stroke engine, that same 1.0 lb-ft of torque per cubic inch is a BMEP of 75.4 psi.

(The discussion on the remainder of this page is with respect to four-stroke engines, but it applies equally to two stroke engines if you simply substitute 75.4 everywhere you see 150.8)

If you know the torque and displacement of an engine, a very practical way to calculate BMEP is:

BMEP (psi) = 150.8 x TORQUE (lb-ft) / DISPLACEMENT (ci)

(Equation 8-a, 4-Stroke Engine)

BMEP (psi) = 75.4 x TORQUE (lb-ft) / DISPLACEMENT (ci)

(Equation 8-b, 2-Stroke Engine)

(IF you prefer pressure readings in Bar rather than PSI, simply divide PSI by 14.5)

(IF you are interested in the derivation of those relationships, it is explained at the bottom of this page.)

This tool is extremely handy to evaluate the performance which is claimed for any particular engine. For example, the 200 HP IO-360 (360 CID) and 300 HP IO-540 (540 CID) Lycomings make their rated power at 2700 RPM. At that RPM, the rated power requires 389 lb-ft and 584 lb-ft of torque respectively. (If you don't understand that calculation, CLICK HERE)

From those torque values, it is easy to see (from Equation 8-a above) that both engines operate at a BMEP of about 163 PSI (12.25 bar, or 1.08 lb-ft of torque per cubic inch) at peak power. The BMEP at peak torque is slightly greater.

For a long-life, naturally-aspirated, SI (spark ignition) gasoline-fueled, two-valve-per-cylinder, pushrod engine, a BMEP over 210 PSI (14.5 bar) is difficult to achieve and requires a serious development program and very specialized components.

It is worthwhile to note that a contemporary, normally-aspirated CI (compression-ignition) engine can easily make 15 bar of BMEP, and several turbocharged CI street engines routinely exceed 20.5 bar. It is helpful to remember that BMEP is a useful tool for comparing and evaluating similar types of engines.

For comparison purposes, let's look at what is commonly believed to be the very pinnacle of engine performance: Formula-1 (Grand Prix).

An F1 engine is purpose-built and essentially unrestricted. For 2006, the rules required a 90° V8 engine of 2.4 liters displacement (146.4 CID) with a maximum bore of 98mm (3.858) and a required bore spacing of 106.5 mm (4.193). The resulting stroke to achieve 2.4 liters is 39.75 mm (1.565) and is implemented with a 180° crankshaft. The typical rod length is approximately 4.016 (102 mm), for a Rod/Stroke ratio of about 2.57. These engines are typically a 4-valve-per cylinder layout with two overhead cams per bank, and pneumatic valvesprings. In addition to the few restrictions stated above, there are the following additional restrictions: (a) no beryllium compounds, (b) no MMC pistons, (c) no variable-length intake pipes, (d) one injector per cylinder, and (e) the requirement that one engine last for two race weekends.

At the end of the 2006 season, most of these F1 engines ran up to 20,000 RPM in a race, and made in the vicinity of 750 HP. One engine for which I have the figures made 755 BHP at an astonishing 19,250 RPM. At a peak power of 755 HP, the torque is 206 lb-ft and peak-power BMEP would be 212 psi. (14.63 bar). Peak torque of 214 lb-ft occurred at 17,000 RPM for a BMEP of 220 psi (15.18 bar). There can be no argument that 212 psi at 19,250 RPM is truly amazing.

However, let's look at some astounding domestic technology.

The NASCAR Cup race engine is a severely-restricted powerplant, allegedly being derived from "production" components, although as of 2010, all 4 engines competing at that level (Chevy, Dodge, Ford, Totota) are purpose-built race engines designed specifically to NASCAR's rule book.

By regulation, they use a cast-iron 90° V8 block and 90° steel crankshaft, with a maximum displacement of 358 CID (5.87 liters). A typical configuration has a 4.185" bore with a 3.25" stroke and a 6.20" conrod (R/S = 1.91). Cylinder heads are similarly purpose-designed and highly-developed, but limited to two valves per cylinder, specific valve angles, specific port floor heights, etc.. The valves are operated by a single, engineblock-mounted, flat-tappet camshaft (that's right, still no rollers as of 2012) and a pushrod / rocker-arm / coil-spring valvetrain. It is further hobbled by the requirement for a single four-barrel carburetor (until 2011) and now (2012 on), by a 4-barrel-carburetor-like throttle body and individial runner EFI. Electronically-controlled ignition is allowed (as of 2012), and there are minimum weight requirements for the conrods and pistons.

How does it perform? In early 2010, the engines were producing in the neighborhood of 860 HP at 9000 RPM (and could produce more at 10,000 RPM, but engine RPM has been restricted by means of a rule limiting the final drive ratio at each venue).

Consider the fact that, to produce 860 HP at 9000 RPM, requires 501 lb-ft of torque, for a peak-power BMEP of nearly 211 PSI (14.55 bar). Peak torque (2010) was typically about 535 lb-ft at 7800 RPM, for a peak BMEP of over 225 psi (15.5 bar).

THAT is truly astonishing. Compare the F1 engine figures to the Cup engine figures for a better grip on just how clever these Cup engine guys are.

That being said, recent winners in the annual Engine Masters competition are achieving over 16.9 bar BMEP (145 psi) with normally-aspirated, petrol-fueled, SI, 2-valve pushrod engines, although the builders freely admit that, due to the very aggressive cam profiles, rocker ratios, gross valve lift numbers, and other compromises aimed at maximizing BMEP, these engines have relatively short life expectancies.

To appreciate the value of this comparison tool, suppose someone offers to sell you a 2.8 liter (171 cubic inch) Ford V6 which allegedly makes 230 HP at 5000 RPM, and is equipped with the standard OEM iron heads and an aftermarket intake manifold and camshaft. You could evaluate the reasonableness of this claim by calculating (a) that 230 HP at 5000 RPM requires 242 lb-ft of torque (230 x 5252 ÷ 5000), and (b) that 242 lb-ft. of torque from 171 cubic inches requires a BMEP of 213 PSI (150.8 x 242 ÷ 171).

You would then dismiss the claim as preposterous because you know that if a guy could do the magic required to make that kind of performance with the stock heads and intake design, he would be renowned as one of the pre-eminent engine gurus in the world. (You would later discover that the engine rating of "230" is actually "Blantonpower", not Horsepower.)

As a matter of fact, in order to get a BMEP value of 214 PSI from our aircraft V8, we had to use extremely well developed, high-flowing, high velocity heads, a specially-developed tuned intake and fuel injection system, very well developed roller-cam profiles and valve train components, and a host of very specialized components which we designed and manufactured.

DERIVATION OF THE BMEP EQUATIONS
The definition of BMEP (Brake Mean Effective Pressure), as previously stated at the top of this page, is: " the mean (average) pressure which, if imposed on the pistons uniformly from the top to the bottom of each power stroke, would produce the measured (brake) power output". AGAIN, NOTE that BMEP is purely theoretical and has nothing to do with actual cylinder pressures.

If we put the definition into mathematical form, we get:,

HP = BMEP x piston area x (stroke / 12) x RPM x power-pulses-per-revolution / 33000

Working through that equation in terms of a single cylinder engine, BMEP (in PSI) multiplied by piston area (square inches) gives the mean force applied to the piston during the power stroke. Multiplying that force by the stroke (inches divided by 12, which changes the units to feet) gives the net WORK (in foot-pounds) produced by the piston moving from TDC to BDC with the BMEP exerted on it throughout that motion. (Clearly this is not an attempt to represent the reality in the combustion chamber. As previously stated, BMEP is simply a convenient tool for comparing and evaluating engine performance.)

Next, power is defined as work-per-unit time. Therefore, multiplying the WORK (ft-lbs) by the RPM, then multiplying by power-pulses-per-revolution (PPR) gives the net (brake) power (foot-pounds per minute in this example) produced by one cylinder. (In a single-cylinder engine, PPR is either 1 for a 2-stroke engine or 1/2 for a 4-stroke engine.

Since one HORSEPOWER is defined as 33,000 foot-pounds-of-work-per-minute, dividing the WORK (ft-lbs) by 33,000 changes the units from foot-pounds-per-minute to HP.

Since it is clear that piston area x stroke is the displacement of one cylinder (in cubic inches), then the equation can be simplified to:

HP = BMEP x (displacement / 12) x RPM x power-pulses-per-revolution / 33000


Horsepower is also defined as:

HP = Torque x RPM / 5252

Substituting that equation into the preceding one gives:

Torque x RPM / 5252 = BMEP x displacement / 12 x RPM x PPR / 33000


Reducing that equation gives:

BMEP = (Torque x 12 x 33,000 / 5252) / (Displacement x PPR)

Evaluating the constants, 12 x 33,000 / 5252 = 75.39985, which can safely be approximated by 75.4. Simplifying the equation again gives:

BMEP = (Torque x 75.4) / (Displacement x PPR)


It is also clear that because the equation includes PPR, it applies to engines with any number of cylinders by using the total displacement, total brake torque, and correct PPR.

Suppose, for example, that you measured 14.45 lb-ft of torque from a 125 cc (7.625 CID) single-cylinder 2-stroke engine at 12,950 RPM, you would have 35.63 HP (285 HP per liter, quite impressive indeed). The BMEP would be:

BMEP = (14.45 x 75.4) / (7.625 x 1) = 142.9 psi (9.85 bar)


That BMEP (9.85 bar) is an impressive number for a piston-ported 2-stroke engine.

However, suppose someone claimed to be making that same torque from a single cylinder 4-stroke 125 cc engine at 12,950 RPM. The power would be the same (35.63 HP, or 285 HP per liter). The power density would not necessarily set off alarms, (the 2008 2.4 liter F1 V8 engines approached 315 HP per liter), but the BMEP would cause that claim to be declared highly questionable:

BMEP = (14.45 x 75.4) / (7.625 x 1/2) = 285.8 psi (19.7 bar)


That BMEP (19.7 bar) is clearly absurd for a normally-aspirated engine. Professor Gordon Blair stated that exceeding 15 bar of BMEP in a N/A engine is virtually impossible, but that was a few years ago. NASCAR Cup engines are now approaching 15.6 bar


Clearly, the difference between 2- and 4-stroke engines is simply a factor of 2, because of the fact that a 2-stroke cylinder fires once per revoultion whereas a 4-stroke engine fires only once per two revolutions. The equations can be simplified further by incorporating that PPR factor in the constant 75.4 and eliminating PPR from the equation, therefore making the constant for a 4-stroke engine 2 x 75.4 = 150.8. That produces the equations shown at the top of this article, which use the full engine displacement and measured torque.

BMEP = 150.8 x TORQUE (lb-ft) / DISPLACEMENT (ci)

(Equation 8-a, 4-Stroke Engine)

BMEP = 75.4 x TORQUE (lb-ft) / DISPLACEMENT (ci)

(Equation 8-b, 2-Stroke Engine)

Happy Birthday Jan!
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Joyeux Anniversaire, Monsieur !!!

Happy birthday Jan.
And congratulations for your numerous descendants.
We wish you many more other happy birthdays.

HAPPY BIRTHDAY, JAN!

happy birthday,
cest pas que le 2 temps que tu est maitre a fabriqué

@David: Your French is certainly better than my English but I think you should write him in English I think it will be easier for him
Voici la traduction de ton texte avec GOOGLE:
"cest pas que le 2 temps que tu est maitre a fabriqué" = this is not the time for you two is master has made
Unbelievable isn't it!

Thanks for the birthday wishes!

HAPPY BIRTHDAY, JAN !

Wat forgotten in this discussion is the rpm's. It is far more difficult to obtain 15 Bar BMEP at 20.000 rpm than it is at 10.000. That is a lot more air!

After all BMEP and rpm's make the power.

Above certain rpm's two valvers get out of breath and four valvers take over to get enough air in the engine.

Also, that is why two strokers are more or less at their limit in the present design. There is simple not enough TA left

Never the less it is very impressive what they acomplish with that old design two valvers.

Frits Overmars a écrit:

Happy Birthday Jan!
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Happy Birthday Jan from one of your extended two stroke family - please know that I appreciate your generous contribution of knowledge and advise on this forum.

Kind regards
Allan.

Frits Overmars a écrit:

Maximum torque of the RSA125 was 29.5 Nm between 12,000 and 12,500 rpm, measured at the secundary gearbox shaft. That should give a BMEP of about 17 bar

Frits, I calculate 14.8 bar for 29.52 Nm and 125 cm³. What am I missing?

Reading glasses?

Frits Overmars a écrit:

Transmission losses are in the region of 5% per transmission step. From the crankshaft to the primary gearbox shaft you loose 5% of the crankshaft power. In the gearbox you loose another 5% of the input.

Frits Overmars a écrit:

....when calculating the mean effective pressure in the cylinder, by definition you must base the calculation on crankshaft power.

See what you missed? The conversion from gearbox shaft torque to crankshaft torque.
The 'about 17 bar' I mentioned is a rounded figure. If you calculate it exactly, with 124,82 cc, it works out at a bit less than that. But it's still more than any naturally aspirated petrol-burning four-stroke will deliver.

Ah, two transmission steps not one. Because one 5% loss would be 15.62 bar and two would make it 16.48 bar MEP if the dyno was braking the crankshaft instead of the transmission output shaft. OK, it makes sense now.

Also, it's not quite 125 cm³ but actually 124.82. That brings it from 16.48 to 16.51 bar. OK, now you can round up to 17.

Brian Callahan a écrit:

Also, it's not quite 125 cm³ but actually 124.82. That brings it from 16.48 to 16.51 bar. OK, now you can round up to 17.

If we really have to go into the small details it must be between 124,89 and 124,91 cm³ becuase the cylinder bore is between 15 and 20 microns over nominal size.

So now you can round down back to 16;)

Frits Overmars a écrit:

Reading glasses?

Hi Fritz,
Thank you for your many informative postings on this forum.
From the many photos of the RSA cylinder on this forum, the exhaust sub ports of this cylinder are interesting. Many of the Japanese GP cylinders have the exhaust sub ports extended down to just above the main transfers with the minimum gap. The RSA is very different in this way. Has the sub port been designed more for the reflected wave from the pipe in mind than total time area for blowdown?

In another page, you mentioned crankcase volume and the effect of crankshaft rotation on the inlet flow; you gave an example of a test by Jan on reversing the crankshaft direction in a kart motor to increase the power - is the inlet flow helped by the crankshaft rotation with the inlet on the back of the RSA cylinder and would it be better if some covering of the crankshaft was present?

Many thanks for your answer in advance.
Regards

GtG001 a écrit:

Hi Fritz Frits,
From the many photos of the RSA cylinder on this forum, the exhaust sub ports of this cylinder are interesting. Many of the Japanese GP cylinders have the exhaust sub ports extended down to just above the main transfers with the minimum gap. The RSA is very different in this way. Has the sub port been designed more for the reflected wave from the pipe in mind than total time area for blowdown?

The amount of material between the auxiliary exhaust ports and the A-transfers in the Aprilia cylinders is minimal too, as you can see in the picture below. Only the rear part of the aux. port bottoms is raised; too close a proximity would enhance short-circuiting via the piston pin bores. Apart from this, the aux. ports have solely been designed for optimum blowdown time.area.
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Citation :

In another page, you mentioned crankcase volume and the effect of crankshaft rotation on the inlet flow; you gave an example of a test by Jan on reversing the crankshaft direction in a kart motor to increase the power - is the inlet flow helped by the crankshaft rotation with the inlet on the back of the RSA cylinder and would it be better if some covering of the crankshaft was present?

Jan's test was not meant to increase the power but merely to establish the influence of the direction of rotation on the inflow, which turned out to be considerable.
Covering the crankshaft is unnecessary when the crankshaft rotates with the flow. When it rotates against the flow, covering is somewhat helpful but it will not cancel the negative effect and it also is disadvantageous to the lubrication of main and big end bearings.

That cylinder has to be close to 100% of bore width in the exhaust port.
I imagine rings were not lasting long at all.
This is a great thread with a whole lot of learning in it

My pleasure, Wax. With you and GtG001 joining in, it is beginning to look like a Skippie invasion here .
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Im a kiwi living in australia , gold coast( bikini central)
I have been building jetski two stroke engines for about 15 years now, Always interesting to read threads like this and get more information. As soon as your to old to learn you may as well curl up and die.