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Any constructive criticism of my updated design approach?
4

Any constructive criticism of my updated design approach?

Any constructive criticism of my updated design approach?

(OP)
All,

I'm now patent pending on my design updates incorporated since filing the original patent, so I can show it now and solicit critique (for those familiar with the system engineering process, the first patent reflected Preliminary Design while the new continuation patent reflects down select of mechanization options to a final design). There are still a number of details to complete, but I estimate final design is about 90% complete at this point.

Folks may recall my engine is a rotating cylinder radial that uses opposed pistons mated to a dedicated charge pump for scavenge/charge and employs a two-stroke Homogeneous Charge Compression Ignition (HCCI) cycle. All pistons are driven by cams, and each set of cylinders completes four full cycles per revolution. With six cylinder sets, the result is 24 complete cycles per revolution. One key difference between the current design and the original is the relocation of the charge pump from its radial position coaxial with the opposed pistons to a new position beside the opposed pistons. This change allows use of a third cam to drive the charge pump piston which previously moved in unison with the outside piston of the opposed pair. This new position shortens the transfer passage between the charge piston and the opposed pair, allows greater flexibility in charge pump timing versus the opposed pair, eliminates the cam shaft (and associated flexure) of the prior design, and maximizes the cam contact area of the heavily loaded pistons of the opposed pair. Combined with a new intake/exhaust port layout, the new approach significantly reduces back pressure (and thus pumping loss) during scavenge/charge. Another key change in the new design is reduction of the innermost piston's stroke to encompass opening and closing the intake port alone while allocating the full compression/expansion stroke to the outer piston of the pair. This change is a simple matter of mechanics; the radius of the inside cam is much smaller than the radius of the outside cam, and the larger radius cams can move further in a given number of degrees and a given material stress limit. The final major difference in the new design is the incorporation of a traditional valve/port controlled version of the Atkinson cycle to mechanize variable compression ratio and control autoignition timing; the inner cam controlling intake port timing is rotated relative to the outer cams so the port remains open during some portion of the compression stroke (note the charge pump cam profile is the inverse of the main piston's cam profile during this period so the net pressure of the Atkinson transfer is near zero). This facility is controlled according to knock sensors in the inner cam as well as atmospheric pressure and temperature sensors to vary compression and ensure ignition occurs at the ideal time in the cycle.

Below is a Solidworks Motion Study animation of one cylinder set (of 6) comprised of an opposed piston pair and a charge pump piston. Note that the animation rotates the cams rather than the pistons and cylinders as in the final design so that the cycle can be better observed. The animation shows the charge pump ports are aligned with the intake ports which are open and closed by the upper (or innermost) piston. Note the circular pocket near the intake ports in the main cylinder; this is where the fuel injectors are installed. Driven by a cam in the side housing, these fuel injectors start injecting at a fixed point in the cycle representing the latest possible closure of the intake port.



Per my math and CAD models, the prototype engine will displace between 25 and 31 cc depending on altitude and ambient conditions (25cc on a standard day at sea level). The engine will be 6 inches in diameter and 5 inches thick and employ a bore of just over 1 inch with stroke sweeping a total of 0.481 inches (including the 0.062 inch tall ports at the top and bottom of the cylinder). It will produce 5.7 HP @ 2,626 RPM (propeller speed) and 11.5 lb-ft of torque with 58.6% efficiency including friction and pumping loss (68.3% theoretical) at sea level. 73% of said performance will be available at 15,000 foot altitude (even though air density is only 62.9% of that at sea level). The weight is high at 16.5 lb, but I expect that to come down to around 10-12 lb after weight reduction is complete (deferred until after all other aspects are proven). Performance figures are, of course, subject to validation in real hardware, but I'm encouraged by the 58.6% efficiency indicated by the models; as long as the end result is above 50%, I've got a real product.

Based on threads regarding the Achates engine, I imagine some will be quick to point out that opposed piston engines tend to dump oil out the cylinder ports. Now that the updated patent is filed I can say that this is one area where a rotating cylinder block is key; the passages to and from the intake and exhaust ports will be tilted slightly inward toward the motor axis such that oil exiting the ports can be collected via centrifugal force and routed back into the low pressure oil loop (which flows nearby through passages surrounding each cylinder for cooling). This of course assumes that the oil exiting the ports is in liquid form, and I'll have to conduct experiments to determine how it exits the port and how best to capture it aided by centrifugal force.

The whole point of this post is to solicit constructive criticism, so don't hold back. I only ask that it be constructive and respectful.

Rod

RE: Any constructive criticism of my updated design approach?

OK, this is the first time I've understood the engine in this long discussion, I guess I'm your visual learner.

Looking at it, I wonder what keeps the pistons from getting cocked and jamming. Which is to say that you've got a moment on the piston which is L*P(time)* piston area(µ + tan(angle of the cam surface)) where L is approx. the distance from the piston ring to the contact point on the cam and P(time) is the pressure at any point in time. Have you looked into this?

Edit: I see the guide now.

I'm still not clear what spins here, the part with the pistons or the cam. Either way, I see the cam vibrating if not well supported.

RE: Any constructive criticism of my updated design approach?

I imagine you need a breather on the guide plate and you'll have an associated pumping loss.

RE: Any constructive criticism of my updated design approach?

(OP)
Moon161,

The odd shape of the inner piston near the inner cam provides a flat surface for transfer of side thrust resulting from cam interaction. The outer piston of the opposed pair and the charge pump piston both have guides as you note in your edit. Your point about needing to vent the guides under the pistons with the long strokes is a good one (thanks!). I'll explore adding holes to the guides or putting another ring of ports around the bottom of those two cylinders for venting. Adding another ring of ports will likely yield the least pumping resistance. I'll have to see if I can route them to the intake channels somehow so I can reuse the intake air filter. Good catch. I can't believe I missed it. That's a good example of the value f peer review!

The cylinder block rotates counterclockwise from the pilot's perspective (per convention) and the cams are stationary. The use of a rotating cylinder block (aka a rotor) facilitates incorporation of a low pressure centrifugal oil pump used for lubrication and cooling as well as the method used to recover oil scraped into the ports.

If I had shown all cylinders operating in the animation, it would be clear that there is always a cylinder set on the opposite side of the rotor that's in precisely the same portion of the cycle. This results in near perfect balance with direct and reaction loads cancelling at the bearings. As for vibration, I believe it will be driven primarily by the firing pattern. If you label the six cylinders North (N), North East (NE), South East (SE), South (S), South West (SW), and North West (NW), the firing pattern is N and S, NE and SW, SE and NW. As each opposing pair fires, it tries to squish the inner cam and stretch the outer cam ever so slightly (making both a bit oval). Fortunately, the outer cam is thickest where the combustion pressure is greatest, and that's where it's attached to the side housings, so there's very little actual movement. Nonetheless, I think the engine will have a characteristic tone at 4x the RPM. I wish I had the processing horsepower to simulate the engine in operation, but I don't; it will have to wait for prototype test.

Rod

RE: Any constructive criticism of my updated design approach?

Also, is it OK for the pistons to float? I don't see any elements that will keep the pistons on the cam except for charge & combustion pressure.

RE: Any constructive criticism of my updated design approach?

Rod,

A couple questions, first, what is forcing the inner pistons to be in contact with and follow the cam? Are there bias/return springs somewhere that aren't shown?

Second, in the below screen grab of your animation it looks like the interface angle between the piston and the cam face on the power stroke is around 45 degrees, meaning that the force rotating the pistons relative to the cam (Fx in the picture) is only about 70% of F neglecting losses to friction. This would seem to hurt your efficiency and create a lot of wear as the piston tries to dig into the cam, how was this angle chosen?

RE: Any constructive criticism of my updated design approach?

(OP)
moon161,

Included in the last 10% of design is the addition of springs. I have the requirements based on the sum of acceleration, gas pressure, and centrifugal force, but haven't yet decided what kind of spring to use. Flat springs are simplest but compression springs are more efficient and don't wear. Just another design decision to work through.

Rod

RE: Any constructive criticism of my updated design approach?

(OP)
hendersdc,

Per my above response to moon161, I'm currently in the process of selecting springs. I have all the forces (requirements), I just haven't decided on which spring to use or how to incorporate them. I don't anticipate problems in this area, but I'm always ready to be surprised at how a minor challenge can become a major challenge (it happens more often than I like to admit).

The maximum pressure angle of the cam you've chosen to illustrate is actually 38.01 degrees (the max pressure angle of the inner cam is 22.12 degrees, and that of the charge pump is 41.82 degrees). Similar to a crankshaft, only a portion of the vertical force applied to the cam is transferred to rotational force. The rest remains unexpended in the combustion gas as potential work (this is why crankshafts are more efficient than some assume based purely on angles). The cam and follower are designed to handle the full maximum force applied at a zero degree pressure angle (unexpected premature detonation) with 100% margin, so there's little or no risk of plastic deformation. Because of the radius of the cam, the surface speeds are high (1509 fpm min, 7927 fpm max), and they combine with the surface width, hardness (aged Maraging 350 steel with 56 Rc), and micro-polished finish to keep the assembly in the hydrodynamic regime at all times. Note, by the way, that there's a 0.039 in fillet on the follower edges that's not shown (I accidentally left them suppressed when I ran the animation).

As for how the contact angles were chosen, I knew I wanted them as large as possible during expansion (not because they're more mechanically efficient but because they reduce the time available for heat transfer and ring blow-by) but also have to keep an eye on acceleration, jerk, and yield margin. I thus created a math model in Excel that calculates total force (acceleration, centrifugal, gas) and yield margin then increased the pressure angle to the maximum possible within the limits dictated by the need to accelerate/decelerate smoothly with minimum jerk while staying withing the stress limits.

The cams are both the boon and the ban of my design. On the plus side, they allow precise control of timing and thus facilitate HCCI. On the down side, they are finicky compared to crankshafts. More time has gone into the cams than any other aspect of the design (with fuel injectors coming in second), and I have a lot of experiments and measurements planned during the prototyping phase. If I succeed, it will be because of the cams. If I fail, it will be because of the cams.

Rod


RE: Any constructive criticism of my updated design approach?

I have never seen any evidence to suggest that cam actuation (ie unusual velocity profiles) is advantageous to initiating HCCI.

je suis charlie

RE: Any constructive criticism of my updated design approach?

I see a mechanism that is rather large relative to the swept volume, allowing for 6 of the piston sets as mentioned.
I don't know much about cams but I'd be interested in the comments of anyone involved in the successful design and release of a camshaft for an engine for commercial application. Particularly concerning the mechanical efficiency, and durability.

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)
Gruntguru,

If you had heard of cam based HCCI using my approach, it would be patented, and my first patent would not have been allowed.

Companies working HCCI seek to incorporate HCCI into standard engines. That doesn’t work well for a variety of reasons. To date I believe only Mazda is committed, and their spark assisted HCCI is not standard HCCI.

My engine consultant worked on several HCCI teams at large auto companies. He was excited after his review of my design and singled out the cams as a new approach worth pursuing because they allow better contol of timing wgich is critical from the perspective of the chemical kinetics. You literally could not attain the timing I’m using with a crankshaft.

Do you any other aspects on you’d like to offer constructive input?

Rod

RE: Any constructive criticism of my updated design approach?

I would suggest some fundamental research on HCCI and dV(olume)/dt profiles. At the very least a single cylinder version of your design.

You may think my criticisms are non-constructive, but my intention is to help you save a lot of money and time.

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)
gruntguru, I have done extensive research into HCCI, and I hired a consultant experienced in HCCI to review my calculations and design. I plan a large number of what we used to call in the defense R&D biz "critical experiments." Every aspect of the engine will be tested and refined in isolation before a complete engine is built. Rod

RE: Any constructive criticism of my updated design approach?

(OP)
hemi,

Quote (hemi)

I see a mechanism that is rather large relative to the swept volume, allowing for 6 of the piston sets as mentioned.

I think you're questioning power density. Below is snapshot of the comparison of my engine to several competing engines. My engine is summarized at top. The GF30, GF38, and G26 are popular engines used in giant-scale model aircraft. The GX25, GX35, and GXH50 are Honda utility engines. I use torque as the basis for comparison reflecting application driving constant speed propellers. Since most competing engines run at high RPM compared to mine, I grant them use of a lossless reduction gear and calculate their torque at 2,626 RPM. Looking at torque alone, the closest competitor has only 60% of my torque. Comparing torque/weight, the model aircraft engines kick my tail but the Honda utility engines aren't even close (note I expect to bring the weight of my engine down substantially by removing unnecessary metal, but I won't put any significant effort into that task until everything is working). Comparing torque/volume, mine wins (this is the basis for my belief mine can be made significantly lighter).



Quote (hemi)

I don't know much about cams but I'd be interested in the comments of anyone involved in the successful design and release of a camshaft for an engine for commercial application. Particularly concerning the mechanical efficiency, and durability.

You're intimately familiar with cams that operate reliably for 100's of thousands of miles. You could imagine my engine using stationary cylinders with the pistons driven by cams on rotating shafts connected by gears. The challenge is friction, and that depends on lubrication. The average outside radius of my outside cam is 2.87 inches and the radius of the follower is .325 inches. Just like a journal bearing, the rotating outer cam is pulling oil under the cam/follower interface. The effectiveness of such bearings is determined by surface speed, interface length, and material hardness. My surface speed is very high (1509 fpm min, 7927 fpm max) as is my hardness (aged Maraging 350 steel with 56 Rc), and the length is substantial (0.989 inches). The trick will be maintaining the proper amount of oil on the cam, and that's a non-trivial challenge. At present, I'm planning to felt wipe the cam surface with oil, but I can't vouch for the efficacy of this approach until I get through the associated critical experiments.

As it stands, my analysis includes the friction of an automotive valve train (the comparison is apt given my "valve" lift is only 0.418 inches and my "valve" mass is only 0.083 lb). Below is a plot of friction losses (Friction Mean Effective Power aka FMEP) by subsystem from Ricardo with a blue line showing total friction according to Heywood. I average the two to estimate FMEP in my analysis. The estimate is, of course, only approximate, so I have to refine it during critical experiments. My plan is to put a heavy spring between the pistons and spin the rotor using an electric motor to estimate friction while measuring motoring torque.

RE: Any constructive criticism of my updated design approach?

Neither torque nor BMEP is a good basis for comparing different engine types targeting the same application unless they happen to have the same output rpm at their respective design points.
You can't use FMEP from a conventional reciprocating engine as an estimate for your FMEP. Nor can you use the [conventional engine's] valvetrain contribution to relative or absolute friction, since the power being dissipated in the valvetrain is but a fraction of the shaft power. You could use the average torque required by a reciprocating engine's valvetrain divided by a summation of the average normal force on the cam lobes as an indicator for the friction torque of your cam system per unit of the summed average normal force. How good of an indicator probably depends on many factors, but it's a starting point.

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

Neither torque nor BMEP is not a good basis for comparing different engine types targeting the same application unless they happen to have the same output rpm at their respective design points

I agree it would be unfair to compare engines designed to operate at different RPMs without gearing them to the application. That's why I granted each engine a "free" (zero size and weight, 100% efficient) transmission to multiply their peak torque (the greater of their stated torque@RPM or their HP*5252/@RPM) and reduce their shaft speed to the common RPM of 2,626.

Quote (hemi)

You can't use FMEP from a conventional reciprocating engine as an estimate for your FMEP

I disagree. It's a reasonable approximation useful until motoring FMEP can be better approximated on a dyno.

Quote (hemi)

Nor can you use the [conventional engine's] valvetrain contribution to relative or absolute friction

I don't use the valvetrain number in isolation. I use the average of the full FMEP described by Heywood's and Ricardo's trendlines of full-engine FMEP over RPM. Of course it's not exact, but the similarity between Heywood's and Ricardo's results indicates there is reasonable correlation of FMEP and RPM. On the plus side, I used the entire estimate including all the components and subsystems that aren't present in my engine.

The fact is, I won't know IMEP, BMEP, FMEP, torque, HP, or efficiency until I test the full engine. My math models only provide approximations to guide design decisions, and I see no alternative to use of informed approximations at this phase in the process. I do plan critical experiments specifically targeting FMEP using a motoring dyno. Even those measurements will be approximations, they'll just be better approximations. The only real measures of performance will be torque, RPM, fuel consumption, and emissions of a full engine running under load.

RE: Any constructive criticism of my updated design approach?

Your piston that operates the piston port (transfer port) needs to close off those ports far enough that the (presumable) piston rings on that piston cover the ports, not just the top of the piston. It'll leak like crazy if only the piston crown (and not the rings) cover the ports.

And on a related note ... Having that piston just cover the ports, while its opposing piston does a full stroke, is thermodynamically identical to having both pistons do a half-stroke and having the combustion chamber in the middle of the cylinder as opposed to at one end with the intake ports barely covered. The thermodynamic cycle only cares that the specified volume (compression) is achieved - it doesn't care how or where. Increasing the stroke of your intake-port-operation piston while reducing the stroke of your main power and exhaust port piston changes nothing in terms of complexity, changes nothing in the thermodyanmics, and will make sure the intake ports are actually covered by the piston rings, and will allow the pressure angle of your power-piston cam-follower against its cam to be made shallower.

I don't see what's driving your intake-port piston to actually go outward and stay in contact with its cam ... especially if you have the cylinder block spinning and the cams stationary. Centrifugal action will drive that piston outward (covering the ports and out of contact with the piston).

Personally? I think FMEP is going to be killer, except maaaaaaybe if all cams have rollerized followers. And if the FMEP isn't killer, leakdown will be. And if that's not killer, heat transfer will be (cylinders too small).

I don't see the advantage of using the cam-follower strategy as opposed to a crank-driven strategy with a roots blower (see Commer Knocker, or Fairbanks-Morse).

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (BrianPeterson)

Your piston that operates the piston port (transfer port) needs to close off those ports far enough that the (presumable) piston rings on that piston cover the ports, not just the top of the piston.
It does cover the port. I'm using a Dykes top ring with the top of the ring level with the crown.

Quote (BrianPeterson)

Having that piston just cover the ports, while its opposing piston does a full stroke, is thermodynamically identical to having both pistons do a half-stroke and having the combustion chamber in the middle of the cylinder as opposed to at one end with the intake ports barely covered. The thermodynamic cycle only cares that the specified volume (compression) is achieved - it doesn't care how or where. Increasing the stroke of your intake-port-operation piston while reducing the stroke of your main power and exhaust port piston changes nothing in terms of complexity, changes nothing in the thermodyanmics, and will make sure the intake ports are actually covered by the piston rings, and will allow the pressure angle of your power-piston cam-follower against its cam to be made shallower.
I originally split the stroke between the inner and outer pistons for the reason you describe. When designing the cam profiles, however, I realised (and observed in results) that the smaller base diameter of the inner cam compared to the outer cam means that a given change in displacement results in a proportionally smaller radius of curvature and that means its stress is higher for a given contact force. As it stands, the inner and outer cams have nearly identical stress margin. Also note I inject fuel from the side at an angle down into the cylinder volume. The injector outlet is just below the intake ports and is slanted down the longest distance to the opposing wall. If I split the stroke between the two pistons, I have to move the injector outlet further down the cylinder wall and have to spray at a shallower angle (the extreme case being a straight shot at the opposing wall). I beleive the slanted spray will have less wall impingement and will yield a better mix.

Quote (BrianPeterson)

I don't see what's driving your intake-port piston to actually go outward and stay in contact with its cam ... especially if you have the cylinder block spinning and the cams stationary. Centrifugal action will drive that piston outward (covering the ports and out of contact with the piston).
Correct. I've already responded to prior comments on the matter explaining that I'm in the process of adding springs to all pistons. Likely flat springs for the inner most piston and compression springs for the others.

Quote (BrianPeterson)

I think FMEP is going to be killer, except maaaaaaybe if all cams have rollerized followers. And if the FMEP isn't killer, leakdown will be. And if that's not killer, heat transfer will be (cylinders too small).
I can't use roller cams; the combination of peak load and surface speeds are too demanding. Not sure why you think leakdown will be a killer. Heat transfer is incorporated in my models and shows I'm losing 213K during expansion from 2150K to 663K. That seems reasonable relative to published papers, and my consultant said it looks about right based on his experience as well.

Quote (BrianPeterson)

I don't see the advantage of using the cam-follower strategy as opposed to a crank-driven strategy with a roots blower (see Commer Knocker, or Fairbanks-Morse).
The cam approach is much smaller than a design using crankshafts, crankshafts don't provide fine control of timing like a cam, and they can't produce four full cycles per revolution. I acknowledge the cams are one of the largest (if not the largest) risk element, and I plan a number of critical experiments focusing on their performance. I'm surprised, however, that you view a cam turning at 2,626 RPM max moving a 0.083 lb piston 0.418 inches over a 20 degree period is neigh onto impossible. Finally, a roots blower would be terribly inefficient compared to my charge pumps (which operate at maximum back pressure of only 0.1 bar).

RE: Any constructive criticism of my updated design approach?

Quote (Rodrico)

agree it would be unfair to compare engines designed to operate at different RPMs without gearing them to the application. That's why I granted each engine a "free" (zero size and weight, 100% efficient) transmission to multiply their peak torque (the greater of their stated torque@RPM or their HP*5252/@RPM) and reduce their shaft speed to the common RPM of 2,626.
Ok, I follow your method now, but it is obscure and needlessly complicated. Engines for the same or similar applications are normally compared via either BMEP or power density (displacement or mass basis). Comparing engines intended for dissimilar applications usually doesn't get you very far due to all the dissimilarities. By dissimilar I also mean physical size of the target application (e.g. toy model airplane vs large scale drone).

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

Quote (RodRico)

gruntguru, I have done extensive research into HCCI
Fundamental research or literature review? What I had in mind - specifically - was a test rig that establishes the benefit or otherwise of novel piston actuation in achieving the elusive goal of stable HCCI over a wide operating range.

My concern is that you are investing a lot of time and effort in detail design prior to any single-cylinder testing of the underlying concept which is critical to success of the entire design.

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (gruntguru)

Fundamental research or literature review? What I had in mind - specifically - was a test rig that establishes the benefit or otherwise of novel piston actuation in achieving the elusive goal of stable HCCI over a wide operating range.
Now I get it. I do plan critical experiments on a test rig and admit that I may have put them off longer than I should. The reason for delay is that I want the cams, pistons, and cooling system to be very close to the final configuration (perhaps that's the difference between "fundamental research" and my "critical experiments"). A common error leading to failed R&D projects (failed meaning they went on longer than they should have) is subtle differences between the critical experiment configuration and the final configuration; while the team thought they had captured every relevant aspect of the design in the experiment, it turned out there were aspects they didn't think of or aspects that changed in the final configuration and altered the expected operation.

In my case, I beleive the cams and fuel injectors are the most critical factors, but the impact of centrifugal force is important as well (it impacts the fuel spray, charge stratification, cylinder cooling, and exhaust scavenging). For that reason, the final experiments will have an opposing pair of cylinder sets in a spinning rotor and stationary cams. I'll be fleshing out the entire list of experiments and designing the associated test rigs and instrumentaton systems shortly after I complete my taxes.

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

Comparing engines intended for dissimilar applications usually doesn't get you very far due to all the dissimilarities. By dissimilar I also mean physical size of the target application (e.g. toy model airplane vs large scale drone).
I think I have been unclear regarding my strategy. The application of the listed engines are identical to mine: model aircraft and lawn tools. Though my math model and the parametric CAD model it drives allow me to explore larger engines, I won't tackle those engines myself. I will focus exclusively on the small 25cc engine, the most difficult both in terms of construction and performance. Assuming the small engine yields competitive advantage in my own testing, I will ship the engine on a small motoring dyno with instrumentation to third party testers, publish results, and seek a partner for the larger engines targeting RQ-7 class drones currently using the AR-741 rotary and derivatives up through light aviation currently using the Rotax 912/14/15 and Continental O-200 class engines.

RE: Any constructive criticism of my updated design approach?

It seems to me you're biting off quite a lot at once. An experienced engine developer would develop and test the novel technologies individually: the cam drive, the fuel injection & combustion recipe, and the cylinder/porting/breathing concept each in individual rig tests. Not cheap or easy, but nothing worthwhile is.

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

It seems to me you're biting off quite a lot at once. An experienced engine developer would develop and test the novel technologies individually: the cam drive, the fuel injection & combustion recipe, and the cylinder/porting/breathing concept each in individual rig tests.
I do plan to test all those subsystems separately. As I noted in a prior reponse, "I'll be fleshing out the entire list of experiments and designing the associated test rigs and instrumentaton systems shortly after I complete my taxes." I do, however, plan on using test rigs constructed of assemblies that are as close to the final configuration as is possible.

RE: Any constructive criticism of my updated design approach?

I see an engine type that will last maybe 24 hours if your lucky, the high contact stresses will be problem. It will be fun to see the video of it running in a few months.
Must be forced induction to make the piston move down?

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (enginesrus)

I see an engine type that will last maybe 24 hours if your lucky, the high contact stresses will be problem
You can tell all that without knowing any of the details? No need to know material, peak load, contact angle, surface speed, finish, hardness, etc?

Quote (enginesrus)

Must be forced induction to make the piston move down?
As noted in several responses above, there will be springs holding pistons against cams. I have the requirements (piston mass, acceleration, centrifugal force, gas pressure), but haven't yet selected the spring type. The engine has a charge pump, but charge pressure before compression in the opposed pair is never above ambient.

Quote (enginesrus)

It will be fun to see the video of it running in a few months
It will be years not months, and it may fail a critical experiment before I get there.

RE: Any constructive criticism of my updated design approach?

Unlike a normal crankshaft reciprocating engine that has a power stroke that lasts over 90 degrees of rotation, this design has about what for the power stroke angle of movement? The piston side loads are likely a deal breaker.

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (enginesrus)

Unlike a normal crankshaft reciprocating engine that has a power stroke that lasts over 90 degrees of rotation, this design has about what for the power stroke angle of movement? The piston side loads are likely a deal breaker.
The power stroke lasts 20 degrees, the max piston force is 2,909 lbf, the max pressure angle is 38 degrees, and the max side force is thus 2293 lbf. Rather than apply side load to piston skirts, it is applied to a guide plate located as close to the cam as possible. The guide plate contact area is .182 in^2 which yields max contact pressure of 12.6 kpsi. The yield spec of Maraging 350 steel @ 500F is 310 kpsi. Where's the problem?

P.S. For completeness, the full power stroke is 0.346 inches, max RPM is 2626, mean piston speed durng expansion is 8.2 m/sec, and piston mass is 0.114 lb.
P.P.S A rapid expansion stroke reduces time available for heat transfer and ring blow-by (albeit at some cost in ring pack friction).

RE: Any constructive criticism of my updated design approach?


Where do the return springs go for the various pistons?

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (BigClive)

Where do the return springs go for the various pistons?
I'm working on putting a clip-on flat spring on the inner pistons and compression springs tucked inside the cam follower of the outside piston pushing against a cup in the guide plate.

RE: Any constructive criticism of my updated design approach?

I assume the output speed is somewhat reduced by your cam arrangement - ie a piston undergoes multiple thermodynamic cycles in one revolution? (A conventional 2 stroke executes 1 thermodynamic cycle per revolution.) Your rpm and torque values would be more interesting if they were corrected for this.

EDIT. Just looked at your animation above and counted four lobes on your cam so 4:1 reduction?

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (gruntguru)

I assume the output speed is somewhat reduced by your cam arrangement - ie a piston undergoes multiple thermodynamic cycles in one revolution?
Yes. I calculate work (Joules) for one cycle of one cylinder every fraction of a degree via pressure * change in volume then sum it all up. From that I calculate IMEP and subtract FMEP to obtain BMEP. I then calculate HP from BMEP using a power-pulses-per-revolution (PPR) value of 24 (6 cylinders x 4 power strokes per revolution).

RE: Any constructive criticism of my updated design approach?

As such, most of the columns in your comparison table above are quite meaningless - especially torque/displacement and torque/mass because you are comparing engines with different "internal gearing".

BMEP (specific torque) and specific power are much more useful for comparing engines - especially ones with completely different architecture.

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (gruntguru)

As such, most of the columns in your comparison table above are quite meaningless - especially torque/displacement and torque/mass because you are comparing engines with different "internal gearing".
If you look closely, you'll see I adjusted everyone's torque to reflect use of a perfect (100% efficient, weightless) reduction gear so all engines could be reasonably compared. I suppose I could have reverse calculated BMEP from their given HP and/or torque specifications, but would that really be any better than simply granting them a perfect reduction gear?

RE: Any constructive criticism of my updated design approach?

That sounds reasonable. I still don't understand the torque/volume column. How is it calculated and what use is it?

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (gruntguru)

I still don't understand the torque/volume column. How is it calculated and what use is it?
It's torque per unit volume (aka "size"). Along with torque per unit weight, it informs me as to the competitiveness of my engine. Torque per unit weight is the more accurate figure because weight is easily measured and routinely reported. The torque by unit volume figure is less solid because, lacking an engine to immerse in water, it requires eyeballing and approximation from provided dimensions and drawings. That being said, it shouldn't be far enough off to explain why the T/in^3 and T/lb figures of the GF30, GF38, and G30 engines are out of whack. I suspect they may be pumping up their HP and torque numbers (model airplane engines are not routinely tested on dynos). As a very last step in the process, I may buy one each of those engines and throw them on my dyno just to get some good compartive data.

RE: Any constructive criticism of my updated design approach?

When calculating torque/mass and torque/vol, did you discount the torque from your motor by 75% to allow for the internal "gearing".

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (gruntguru)

When calculating torque/mass and torque/vol, did you discount the torque from your motor by 75% to allow for the internal "gearing".

I don't understand why you suggest I "discount the torque... by 75% to allow for internal gearing." The negative effect of my "internal gearing" is manifest in my FMEP calculations which average the Heywood and Ricardo estimates at 4x my RMP (2,626 x 4 = 10,504 RPM). As I have previously mentioned, this is only an estimate, but I consider it reasonable since Ricardo and Heywood both include sources of friction that are absent in my design. I plan to refine the estimate with measurements early in the critical experiment phase.

RE: Any constructive criticism of my updated design approach?

Sorry, I just noticed you accounted for the "gearing" in the torque figures for the comparison engines.

je suis charlie

RE: Any constructive criticism of my updated design approach?

That's why it's best to stick with conventional measures for comparing engines, i.e. BMEP and power density. Lest you confuse us old curmudgeoneers.

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

What are all the reciprocating parts made of? And piston follower guide material? Any pressure lube to any parts?

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

That's why it's best to stick with conventional measures for comparing engines, i.e. BMEP and power density. Lest you confuse us old curmudgeoneers.

Here you go. The equation deriving BMEP shown in blue is from EPI Inc's wonderful web site. Note that "PPR" is Power Pulses per Revolution which relates to the engine type (4 stroke PPR = 0.5, 2 Stroke = 1, Mine = 4).



Quote (enginesrus)

What are all the reciprocating parts made of? And piston follower guide material? Any pressure lube to any parts?

All recipricating components as well as the cylinder liners, cams, and piston guides are designed assuming use of aged Maraging 350 steel. The main bearings are deep groove roller bearings, all other friction is sliding. Cam surface speeds are high, so I expect them to be in the hydrodynamic regime. The piston guides (essentially piston skirts) and ring pack are similar to a traditional engine. Other then the pressure developed by hydrodynamic operation of the cams, all lubrication is low pressure.

RE: Any constructive criticism of my updated design approach?

To be clear, displacement needs to be figured as swept volume per revolution. Is that what you did?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

To be clear, displacement needs to be figured as swept volume per revolution. Is that what you did?

Look at the denominator of the equation below the table. My engine's Power Pulses per Revolution (PPR) is 4 while that of the others is only 0.5 (4 stroke) or 1 (2 stroke).

RE: Any constructive criticism of my updated design approach?

Doesn't answer my question. What is the swept volume per revolution?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

Doesn't answer my question. What is the swept volume per revolution?

I did answer your question by showing that the swept volume per revolution is inherent to the term "Displacement x PPR" in the BMEP equation. If you're interested, each cylinder displaces 4.17 cc and there are six, so the combined displacement is 25cc. These cylinders complete four cycles per revolution of the output shaft which turns at 2,626 RPM max. As I learned from gruntguru some time back, I could just as easily say my engine is a two-stroke with one quarter the torque operating at 10,504 RPM with a 4:1 reduction gear on the output shaft. Based on my cam loading analysis, I could in fact design the engine to have one power stroke per revolution and operate at 10,504 RPM, but then I'd need a 4:1 reduction gear to bring the shaft speed down to that associated with efficient propellers. Thus, the four power strokes per revolution is simply a more efficient way to mechanize a 4:1 reduction gear.

RE: Any constructive criticism of my updated design approach?

If one cylinder displaces 4.17 cc in 1 up and down cycle, there are 6 of these cylinders, and there are 4 up-down cycles per revolution, then your swept volume per revolution is 100cc. With that swept volume per revolution I get BMEP=36.4
To be clear, only power producing cylinders should be counted in the swept volume. Are we on the same page there?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

Are we on the same page there?

No.

From the "Derivation of the BMEP Equations" at the bottom of the page at www.epi-eng.com/piston_engine_technology/bmep_perf...,
BMEP = (Torque * 12 * 33,000/5252)/(displacement * PPR) where BMEP is in psi, Torque is in lb-ft, displacement is in in^3, and PPR is the number of power pulses per revolution.

Your "up-down cycles per revolution" language is already inherent to the denominator of the equation in the form of power pulses per revolution.

The displacement of the engine is 1.526 ci and the engine creates four power pulses per revolution, so the denominator of the BPEM equation is 6.104 (your 100cc).
The torque of my engine is 11.563 lb-ft so the numerator of the equation is 11.563 * 12 * 33,000/5252 or 871.848, so BMEP = 871.848/6.104 or 142.8 psi.

If you insist my displacement is 100cc (1.526 in^3), you should use the equations at the bottom of the "Derivation of the BMEP Equations" section.
For a two-stroke engine the equation is BMEP = 75.4 * Torque / Displacement. Plugging in my numbers, BMEP = 75.4 * Torque/Displacement = 75.4 * 11.563 / 1.526 = 142.8 psi.

I've posted my equations and their reference source. Please post your equations and reference source so we can compare.

P.S. Note my thermodynamic model calculates Work, IMEP, FMEP, and BMEP from the bottom up then calculates HP and Torque. If you want to go that route, the key figures for one cycle of one cylinder at 2,626 RPM are Work = 4.3 Joules, Displacement = 4.17E-06 m^3, IMEP = 10.4 bar, FMEP = 0.53 Bar, and the engine completes 24 full cycles per revolution.

RE: Any constructive criticism of my updated design approach?

In one revolution, how much volume does your engine displace?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

In one revolution, how much volume does your engine displace?

That's not the standard definition used when labeling an engine's displacement. Most folks just take the swept+clearance volume of one cylinder and mutiply it by the number of cylinders. This is evident in how we calculate the displacement of two-stroke and four-stroke engines; two engines of the same bore, stroke, and clearence volume are labeled as having the same displacement even if one of them is two-stroke and the other is a four-stroke. This in spite of the fact the two-stroke processes twice as much air through combustion every revolution as a four-stroke.

By the standard convention relating Power Strokes per Revolution (PPR) to the stroke class of an engine (Stroke Class = 2/PPR), mine happens to be "half-stroke." It shouldn't be penalized relative to a two-stroke by quadrupling its displacement any more than a two-stroke is penalized relative to a four-stroke by doubling its displacement. For reference, here are the standard equations showing the relation between stroke-class and Power Pulses per Revolution (PPR) in BMEP calculations:

Generic    BMEP = (Torque x 12 x 33,000 / 5252) / (Displacement x PPR)
4.0-Stroke BMEP = (Torque x 12 x 33,000 / 5252) / (Displacement x 0.5) = 150.80 x Torque/Displacement
2.0-Stroke BMEP = (Torque x 12 x 33,000 / 5252) / (Displacement x 1.0) =  75.40 x Torque/Displacement
0.5-Stroke BMEP = (Torque x 12 x 33,000 / 5252) / (Displacement x 4.0) =  18.85 x Torque/Displacement
 
Another perspective is that my engine processes the same volume of air per unit of time through combustion as a 50cc two-stroke running at 10,504 RPM. Nobody drives a car wheel or weed whacker at 10,504 RPM, they employ some form of speed reduction between the engine and the working shaft. My engine just happens to have a clever 4:1 reduction facility integral to the design. This perspective is bolstered by my piston speed; each power stroke completes in only 23.6 degrees of output shaft rotation.

All of the above is symantics, so one could ask why I care. Simple. Many countries and states limit moped displacement to 50cc. Doubling the displacement of a two-stroke because it processes twice as much air through combustion as a four stroke every revolution would severely limit the attractiveness of two-stroke moped engines. Likewise, quadrupling the displacement of my engine because it processes four times as much air per revolution as a two-stroke would limit the attractiveness of my engine. There are similar displacement rules establishing different classes of engines for emissions requirements. It does matter.

Determining the legal displacement of engines that don't use a crankshaft has been painful in the past (some of us no doubt recall the debate regarding the displacement of Mazda's Wankels). Setting the above semantics of displacement aside, my real concern is that you came up with an absurdly low figure for BMEP (36.4 psi) relative to mine (142 psi). I've asked you to post your equations and references so I can compare, but you've yet to do so.

RE: Any constructive criticism of my updated design approach?

(11.6*12*33,000/5252)/(6*4) = 36.4

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

BMEP aside, I can't recall that you ever showed us a PV diagram. Do you have one?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

(11.6*12*33,000/5252)/(6*4) = 36.4

You've double-booked a factor of four in the denominator.
The displacement of my engine independent of PPR is 25cc. You made it 100cc by multiplying by PPR, so you don't need to multiply by PPR a second time.
 
If you want to use the 100cc perspective, then displacement is 6.10 in^3 and PPR is 1, so the denominator should be (6.10*1) 
If you want to use the  25cc perspective, then displacement is 1.53 in^3 and PPR is 4, so the denominator should be (1.53*4)
 
Both denominators are the same and yield the same BMEP, but saying my engine has only one Power Pulse per Revolution is incorrect. It has four.

I'll plot PV tomorrow.

RE: Any constructive criticism of my updated design approach?

To me, and now with the benefit of seeing the animation, the idealised PV diagram will look like that of an idealised two-stroke Otto cycle. (Constant-volume combustion at the moment of minimum volume.) Of course, the real-world combustion process will differ from the ideal. Scavenging, in the ideal cycle, happens at constant (atmospheric) pressure. Of course, there will be pressure losses in real-world operation.

RE: Any constructive criticism of my updated design approach?

(OP)
Here are plots of cam profiles as well as the PV diagram of one opposed piston pair at sea level with Atkinson mode active (displacement|compression ratio reduced from max 31cc|36:1 to 25cc|29:1 via late intake port closure). Several attributes of the cam and port timing are worthy of description. Note that scavenge starts and ends with both the exhaust and intake ports wide open (no gas transfer with partial port openings), and the full volume of air is replaced at atmospheric pressure. Compression doesn't start until the intake port closes a little late in this case which indicates the Atkinson mode is active. After compression, the cylinder is held at minimum volume for a short while to allow completion of the rapid HCCI combustion event at constant volume. Combustion gasses are then expanded at a constant ratio of 1:36. Next, the exhaust port opens to start blow-down. Once blow-down completes, the intake port opens and the cycle begins anew.



The primary purpose of the Atkinson mode is to allow control over compression ratio which is modified according to ambient air temperature and pressure to improve performance at altitude and provide the extra compression needed for cold start at sea level. The variable compression capability is also used to control autoignition timing informed by knock sensors. Atkinson operation is controlled by modifying intake port closure timing relative to exhaust port closure at the start of compression. This is accomplished by rotating the inner cam relative to the outer cam using a servo.

Note the PV plot does not include the Charge Pump because it's a bit confusing when put on the same plot as the opposed piston PV. The volume of the charge pump and the main cylinder ranges from 2x the main cylinder volume plus manifold volume down to 1x the main cylinder plus manifold volume. The cylinders have 0.56 cm^2 of port area servicing 5.2 cc of cylinder volume and gas transport during scavenge/charge is rate controlled by design to yield no more than 0.1 bar back pressure/vaccum.

RE: Any constructive criticism of my updated design approach?

Quote (RodRico)

You've double-booked a factor of four in the denominator.
The displacement of my engine independent of PPR is 25cc. You made it 100cc by multiplying by PPR, so you don't need to multiply by PPR a second time.
I stand corrected. The 4 lobe cam had me barking up the wrong tree.

What IMEP do you get from your PV curve?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

Dwell at TDC would be unnecessary under HCCI (very rapid combustion) conditions and is undesirable due to increased heat loss. Also there are difficulties reducing the high "jerk" arising from a "flat topped" velocity profile for the piston.

It appears that you no longer expect the secondary piston motion to initiate HCCI - congrats.

Using secondary piston phasing to vary CR is a strong feature of your design.

je suis charlie

RE: Any constructive criticism of my updated design approach?

I'm no expert on cam profiles, but unless I'm missing something the ramp accelerations look pretty severe. What's the story?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

Probably OK while it is still "on paper". Just investigating gas exchange and thermodynamics.

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

What IMEP do you get from your PV curve?
For one cylinder executing one cycle, displacement = 4.167E-06 m^3, work = 4.32 J, IMEP = 10.38 bar, FMEP = 0.53 bar, and BMEP = 9.85 bar.

Quote (gruntguru)

Dwell at TDC would be unnecessary under HCCI (very rapid combustion) conditions and is undesirable due to increased heat loss.
The time duration of the flat at max RPM is equal to the fuel's ignition delay at the peak compression temperature. The knock sensors and variable CR will be used to position the combustion event at the end of the flat where expansion begins so heat and pressure loss are minimized. Setting the flat to full ignition delay may be overly conservative. I will adjust or eliminate the flat based on measured data during testing.

Quote (gruntguru)

There are difficulties reducing the high "jerk" arising from a "flat topped" velocity profile for the piston.
True, but it has to stay until I know I don't need it.

Quote (gruntguru)

Using secondary piston phasing to vary CR is a strong feature of your design.
I agree. Other features highlighted in the patent include near uniform cylinder temperature due to the radial cylinder arrangement (eliminating the "cold" cylinders at the ends of an inline or V arrangement that wreck havoc on HCCI control), the integration of a 4:1 propeller reduction gear via four cycles per revolution, the "free" and highly reliable oil/cooling pump provided by the spinning rotor, and use of centrifugal force to recover oil scraped into the ports.

Quote (hemi)

The ramp accelerations look pretty severe. What's the story?
The plots I provided didn't show the cam profile (which defines acceleration and minimum follower radius), they showed piston lift resulting from the combined effect of the cam profile and follower pressure angle (which is useful for visualizing port timing). In the plots below, you can see how the piston lift plot tends to sharpen some transitions and soften others compared to the naked cam profile. Regardless, both plots show high acceleration. My compression ratio at sea level standard conditions is 29:1 which is much higher than normal for an HCCI engine and is only possible because I get through the compression stroke faster than the fuel's ignition delay even at minimum RPM. Rapid stroke is also desireable on both compression and expansion to minimize blow-by and heat transfer. Note the pump piston's acceleration is the highest of the three, but that piston is never subjected to compression or combustion, is significantly lighter, and thus has less cam stress even at higher acceleration. Ultimately, the limit is defined by material yield limits, and all cams operate with significant margin.


RE: Any constructive criticism of my updated design approach?

Your clarification of cam profile accepted for now. When it comes to cutting metal, be prepared for a learning curve on what is feasible to achieve a durable cam system.

Quote (RodRico)

For one cylinder executing one cycle, displacement = 4.167E-06 m^3, work = 4.32 J, IMEP = 10.38 bar, FMEP = 0.53 bar, and BMEP = 9.85 bar.
where is PMEP in all that?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

Be prepared for a learning curve on what is feasible to achieve a durable cam system.
I am fully prepared to have to mess with the cam profiles during test. I am not as pessimistic as you, however, as I am only moving 73.5 grams (before I've put effort into lightening pistons) over 10.4 mm in 1.5 ms. These figures aren't unlike those of modern high performance engines. My cam performance curves (speed, acceleration, and jerk) are all in units of seconds right now but I will soon convert them to units of degrees so they can be better compared to existing cam designs.

Quote (hemi)

Where is PMEP in all that?
PMEP is rolled up in work. If I take it out of work and treat it as a value subtracted from IMEP, the affected figures would be: Work 5.16 J, IMEP 12.38 bar, PMEP 2.01 bar, FMEP 0.53 bar, BMEP 9.85 bar. My model actually works by calculating the port area required to move the desired air mass with a specified pressure differential (0.1 bar) over the specified time allocated for gas exchange in the cam profiles (17 degrees at 2,626 RPM). As I increase the pressure differential, the port size decreases, the cam requirement becomes less demanding (less distance traveled in 17 degrees), and the engine diameter decreases by roughly 2x the resulting change in port height.

RE: Any constructive criticism of my updated design approach?

And this PMEP includes the negative work in the working cylinder as well as the net work in the pumping cylinder, correct?

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

This PMEP includes the negative work in the working cylinder as well as the net work in the pumping cylinder, correct?
Yes, and it also includes the intake manifold, the manifold between the charge pump and working cylinder, and the exhaust manifold.

EDIT: The one pumping loss that's not accounted for is that associated with the Atkinson mode when compression continues as the intake port closes.

RE: Any constructive criticism of my updated design approach?

Rodrico, have you ever looked at pattakon's stuff?

While the actuation is completely different this motor from their website has a similar concept of one piston charging another
https://www.pattakon.com/pattakonOPRE.htm

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (moon161)

Have you ever looked at pattakon's stuff?

Yes, I ran into his material during patent search for my first engine patent. He's using the back face of the opposed pistons to charge the opposed pair. This approach was, I believe, first put forward by an English company called Doxford back when opposed piston engines were all the rage. I used a similar approach in my first engine where the charge pump piston was coaxial with and moved in unison with the outer piston of the opposed pair. To a large extend these approaches using the bottom of a piston to provide the intake charge isn't unlike a standard two-stroke. It has problems, however.

I transitioned to an independently controlled charge piston because the prior approach wasn't delivering the air at the right time. As a result, I was pressurizing teh charge air (not unlike a traditional two stroke) when there is no need. The new arrangement allows me to deliver fresh air to the opposed pistons when needed with only 0.1 bar back pressure, and this yields a huge reduction in pumping loss.

It's important to note that my patents both apply to the class of opposed-piston rotating cylinder engines. I use the centrifugal force of the spinning cylinder block for two primary functions: 1) Oil scraped out of the intake and exhaust ports (a common problem in opposed piston two strokes) is collected and returned to the oil loop via centrifugal force; and 2) a zero-component centrifugal oil pump distributes oil throughout the block for the purpose of cooling and lubrication.

RE: Any constructive criticism of my updated design approach?

(OP)
Well, my first patent for my engine has been granted and published. Skimming through it, I'm struck by how many refinements have been made since it was filed. Many of them were inspired by constructive criticism offered by members of this forum while others resulted from input by my consultant and personal reflection:
  • Outer cam follower shafts eliminated in favor of direct contact of piston with cam. This eliminates shaft flexure, maximizes follower radius for reduced cam stress, and reduces weight.
  • Air pump pistons on their own cam to allow optimized timing. This allows scavenge/charge air to be moved only when the intake and exhaust ports are fully open, thus reducing back pressure and pumping loss.
  • Four to one increase in port area and transfer manifold volume to reduce pumping loss.
  • Multiple reed valves per cylinder set eliminated via optimized gas exchange timing and a simple rotary port. This reduces cost, complexity, and pumping loss.
  • Fuel injectors moved to rotor to spray directly into combustion chamber. This reduces manifold complexity and volume as well as pumping loss while improving mixing.
  • Reduction of piston sets from 12 to 6 which reduces complexity and facilitates direct injection (which adds a fuel injector to each cylinder set).
  • Use of a common bore in the opposed pistons which reduces stress and simplifies construction.
  • Use of inner piston to service intake port only (no role in compression). This maximizes inner cam minimum radius, reduces cam stress, and facilitates Atkinson mode (delayed intake closure).
  • Atkinson mode incorporated for HCCI control to replace multiple control levers. This reduces component count and simplifies control over a wide range of ambient pressure and temperature.
  • Switched from air to oil cooling. This eliminates the need for both oil and air pumping and cooling.
  • Peak combustion pressure reduced from unrealistically high values to 220 bar max. This eliminates risk of Low Speed Premature Ignition (LSPI, autoignition of lubricating oil)
  • Peak combustion temperature below 2150K at all times. This reduces NOX production allowing operation without an NOX adsorber. Note HCCI's inherently low soot production eliminates the need for a Diesel Particulate Filter (DPF), so all that remains is a simple two-way (CO, HC) catalytic converter.
A continuation patent covering all these refinements has been filed, so the refined engine is patent pending. I'm no longer working architectural refinements but design details such as incorporation of piston springs (which is turning ut to be harder to do well than expected), tolerancing, and stack-up.

I know I often seems resistant to criticism, but I do hear you, and am grateful for your contributions.

Thank you all!

RE: Any constructive criticism of my updated design approach?

The patent files are littered with engine concepts that have never been built and run. So, prove us wrong - build it and run it!

"Schiefgehen wird, was schiefgehen kann" - das Murphygesetz

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (hemi)

prove us wrong - build it and run it!

I’m working on it. These things take a remarkable amount of time and effort. I’m used to having a large team moving things along more quickly, but that’s not financially realistic without investors, and I’ll not solicit investors until I have encouraging test results.

RE: Any constructive criticism of my updated design approach?

if when you solicit investors be very careful. you could run into a conman who said he had 2 mill ready to invest but in reality nothing but you are the bait for his personal runaway ponzie scheme so attracts other investors and pays himself 100K a year and minimal on development, marketing . in fact works against development in order to tell investors we need more money. you resign from the company and he takes you to court with a mountain of false paperwork and a big name fancy barrister .The judge sees through the false accusations but tries to protect the innocent shareholders. and awards costs to you and refers the case to ASIC to investigate . expecting only one honest man left standing naturally the conman does not pay costs. and ASIC does nothing because it falls under their threshold of 2 mill a year . The conman takes out a nothing patent in his own name as inventor in order to continue.
the truth is my shield

A tidy mind not intelligent as it ignors the random opportunities of total chaos. Thats my excuse anyway
Malbeare
www.sixstroke.com

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (malbeare)

If when you solicit investors be very careful. You could run into a conman who said he had 2 mill ready to invest but in reality nothing but you are the bait for his personal runaway ponzie scheme so attracts other investors and pays himself 100K a year and minimal on development, marketing.

I already ran into such a fellow. He has a history of working on new engine startups then moving on to the next just as the prior one dies for lack of funding. His primary qualification per his own statements is that he knows how to raise money. Given the trail of failures behind him, I doubt the investment community he works with is excited to see him at their door.

RE: Any constructive criticism of my updated design approach?

The vast majority of novel engine startups fail. This guy would be typical of anyone that specialises in capital raising for such startups.

je suis charlie

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (gruntguru)

The vast majority of novel engine startups fail. This guy would be typical of anyone that specialises in capital raising for such startups.

True dat. Good point.

RE: Any constructive criticism of my updated design approach?

Ditch the OD (bottom) cam and go desmodronic, ducati style. A V8 used a sumilar style once. No more worry about return springs.

How much HP is lost to the valve train? I don't really know but I would think you could gain some efficiency removing the spring.

Also it seems like you could greately simplify this and build a single combustion chamber prototype with off the shelf parts.

Or were you relying on "centrifugal" force to do something?

Here's a video you may find entertaining :)

https://www.youtube.com/watch?v=mkQ2pXkYjRM

Engineering student. Electrical or mechanical, I can't decide!
Minoring in psychology

RE: Any constructive criticism of my updated design approach?

Beautiful drawings.

Where is the cooling system?

Mike Halloran
Corinth, NY, USA

RE: Any constructive criticism of my updated design approach?

(OP)

Quote (michaelwoodcoc)

Ditch the OD (bottom) cam and go desmodronic, ducati style
There's very little load associated with pulling either piston, so I'm considering extending the cam follower with small pins that ride in a slot to either side.

Quote (michaelwoodcoc)

How much HP is lost to the valve train?
It's a piston ported two-stroke, so there's no valve train per se. The cams moving small pistons is somewhat similar, however. I don't know how much HP they're costing, but I do use the Heywood/Ricardo estimates for total engine friction (FMEP). At present, it FMEP is estimated at 0.53 bar per cylinder. If I set FMEP to zero, I see a 0.3 increase in HP.

Quote (michaelwoodcoc)

Also it seems like you could greately simplify this and build a single combustion chamber prototype with off the shelf parts.
Think of the inner piston as an intake valve; its stroke is limited to that required to open and close the intake ports. Using piston gated ports yields greater port area, less pumping loss, reduced surface area of the combustion chamber, and reduced heat loss.

Quote (michaelwoodcoc)

Or were you relying on "centrifugal" force to do something?
Centrifugal force is used to pump oil for cooling and lubrication and to trap/recover oil escaping out the intake and exhaust ports (a common problem with opposed piston two-strokes).

RE: Any constructive criticism of my updated design approach?

Sorry RodRico, I didn't make myself sufficiently clear on any of my points!

I was just thinking out loud on the valve train. In a conventional engine, I bet very little of the energy used to open the valves (if any) is returned. At lower RPM's turning an engine over by hand it is possible for the valve springs to act on the cam and "cog" it to another spot, for lack of a better term, but I would think at higher RPM's the valve train is mostly just a loss.

I brought that up simply because of the mention of return springs for the pistons IIRC.

Here's the valve return system on the V8 that I was talking about: https://www.youtube.com/watch?v=GXIdOTkJhEY

I imagine if you did a groove you could use roller/ball bearings to actuate/move the piston in both directions and use a much more easy to machine metal for a prototype, so even if you just do it to test the engine, can always switch back later.

On the combustion chamber part I think I didn't make this sufficiently clear either. I understand your intention of the inner piston. I think it may be prudent to build a prototype "single combustion cylinder" engine which would really have two cylinders.

The video I posted could be inspiration for a way to design a much more simple single cylinder prototype. There has to be an infinite number of ways that you could use to obtain the same piston movement profile as you would obtain with that "cam" setup.
Who says it has to be cams for the prototype? Sure, losses due to friction would be totally different. But you could calculate that for both setups and see how it would compare.

The oil separation idea seems like a good one. I'd be curious to see itin action.

Engineering student. Electrical or mechanical, I can't decide!
Minoring in psychology

RE: Any constructive criticism of my updated design approach?

@hemi

""It seems to me you're biting off quite a lot at once. An experienced engine developer would develop and test the novel technologies individually: the cam drive, the fuel injection & combustion recipe, and the cylinder/porting/breathing concept each in individual rig tests. Not cheap or easy, but nothing worthwhile is.""

Don't mean to hijack the thread but your post reminded me of Dan Gurney's 1966 quote, "We had a very successful year in terms of mechanical failures."

This was because they went ahead and built the Westlake V12 with surplus WWI machine tools. Of course that mistake led to the Gurney-Westlake Eagle engines, MK1-F1 and MK2-Indy and the rest is history.

RE: Any constructive criticism of my updated design approach?

Quote (michaelwodcoc)

Here's the valve return system on the V8 that I was talking about: https://www.youtube.com/watch?v=GXIdOTkJhEY
A desmodromic pushrod engine! What could possibly go wrong?

je suis charlie

RE: Any constructive criticism of my updated design approach?

LOL, 62.5 lbs. of oil pressure and one drop of oil that isn't smoke in the the video.

RE: Any constructive criticism of my updated design approach?

Yeah I totally don't like what they did in that video, but the theory is what matters I guess.

Here's it done more properly: https://www.cycleworld.com/2014/04/25/cw-tech-valv...



Ducati hasn't every been the front runner for very long in MotoGP. They have quite a sporadic history.

But! They were the only manufacturer that was able to release a version of their MOTOGP bike with the same valvetrain to the public back then in what was it, the early 2000's?

While the big 4 in moto GP all seemed to be using pneumatic valves, Ducati was desmodronic, as they call it.

Pneumatic springs are cool, but then you need a pressure source. IIRC the motoGP bikes used high pressure nitrogen? I remember I was pretty young back then, when I first saw the small tank, thinking to myself: "They're cheating! They have nitrous!" it was then that I first read about pneumatic valves.

Anyways, with the right design, I think you will find desmodronic valves very reliable.

Engineering student. Electrical or mechanical, I can't decide!
Minoring in psychology

RE: Any constructive criticism of my updated design approach?

(OP)
Michaelwoodcoc, The Ducati system is too complex for my purposes (I seek low cost, light weight, and small size). You mentioned you doubt much of the valve spring energy is recovered in a conventional engine, but I assure you it is definately recovered in mine. Remember, the large outer cam of my engine is where linear piston motion is converted to rotary motion of the rotor and output shaft. The work put into compressing the spring during the compression stroke is added to expansion work during the power stroke. Assuming use of compression springs (rather than flat springs which suffer some additional friction loss), the only loss is in a small amount of heating due to material stress in the spring.

RE: Any constructive criticism of my updated design approach?

I am not very good at communicating without pictures, just imagine if:


There were a slot, there would be two sides, ofcourse, and a pin would ride in the slot. The movement of the piston could be controlled in both directions without a spring.

You could do this for every piston.

Because what about the top piston? Wouldn't there be centripital acceleration forcing it down, and where would be the return spring?

I imagine you could have a slot milled and have a rolling bearing riding in the slot if you wished. It'd have to be multiple at that point.

I just imagine your "cam" profile is probably a ground surface. It may be much easier to make a prototype with a slot and bearings, ignoring longevity.

I guess I'm not criticizing at all, just imagining how it could be made in a home garage, more or less, with a 2.5 axis machine you can find used for a couple hundred bucks. (is that the term for machines that can only move the x and y at the same time?)

Engineering student. Electrical or mechanical, I can't decide!
Minoring in psychology

RE: Any constructive criticism of my updated design approach?

(OP)
michaelwoodcoc, Your illustration shows a slotted cam (albeit one with an extreme profile) which is what I mentioned in earlier comments. A few things to note, however. First, return force is far less than driving force, so the retracting pins need not have the same large radius as the driving face, and making them smaller decreases overal engine diameter. Second, a slotted cam still needs spring loading or it will chatter during direction reversal. This loading, however, is much easier to provide as the end-stroke loads are low and spring travel need only encompass the cam slot tolerance (~0.005 to 0.010 inches).

As for roller bearings... Even in the smallest version of my engine, the product of surface speed and load exceeds roller bearing capability, and the problem gets worse as the engine size increases. In addition, use of such bearings increases cost, complexity, and reciprocating mass. I decided to keep it cheap, simple, and low mass by exploiting the high surface speeds to keep a simple sliding cam operating in the hydrodynamic regime from idle to max RPM.

Yes, cams will be made of very hard (Maraging 350) steel machined, ground, and highly polished to optimize lubrication. I may even send the final versions out for diamond coating. Note that cam subsystem performance and life will be the focus of independent subassembly tests.

I am using a Microproto VSS MicroMill DSLS 3000AB equipped with the 4th axis and a Microproto VSS TaigTurn 2000AB for prototype work. I'm starting with a model machined from clear acrylic to check out CNC code and confirm everything fits properly. I will then *attempt* metal components for subsystem test. If I am unable to get the accuracy and surface finish I require on some components, said components will be sent out to a machine shop with the required capability. If nothing else, I figure my acrylic model will ensure the engine is machinable and that everything fits properly so I can minimize the number of costly iterations I go through at an external machine shop.

RE: Any constructive criticism of my updated design approach?

That's quite the home setup, sounds like you don't need to do an easier to machine proto first. I like the idea of the acrylic model. Can't wait to see it, if you're kind enough to share ofcourse!

may be worth considering potential failures though with piston return springs. (Do I remember that correctly?)

What happens if the engine has been sitting up for a long time, and there's a minute amount of corossion in there? Are the springs going to be strong enough? Assuming the engine was stopped with the pistons all at "TDC. Sometimes rings get "stuck". Who knows. But in a typical engine, a rotating crankshaft is able to man handle the piston through these instances that create strong static friction.

Engineering student. Electrical or mechanical, I can't decide!
Minoring in psychology

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