The Rotating Liner Engine (RLE) Diesel Prototype: Reducing Internal Engine Friction by About 40% Under Idle Conditions

applied sciences

Article The Rotating Liner Engine (RLE) Diesel Prototype: Reducing Internal Engine Friction by about 40% under Idle Conditions

Dimitrios Dardalis 1,*, Amiyo Basu 2,*, Matt J. Hall 2,* and Ronald D. Mattthews 2,*

1 RSET Inc., University of Texas at Austin, Austin, TX 78712, USA 2 University of Texas at Austin, Austin, TX 78712, USA * Correspondence: [email protected] (D.D.); [email protected] (A.B.); [email protected] (M.J.H.); [email protected] (R.D.M.)

Abstract: The Rotating Liner Engine (RLE) concept is a design concept for internal combustion engines, where the cylinder liner rotates at a surface speed of 2–4 m/s in order to assist piston ring lubrication. Specifically, we have evidence from prior art and from our own research that the above rotation has the potential to eliminate the metal-to-metal contact/boundary friction that exists close to the piston reversal areas. This frictional source becomes a significant energy loss, especially in the compression/expansion part of the cycle, when the gas pressure that loads the piston rings and skirts is high. This paper describes the Diesel RLE prototype constructed from a Cummins 4BT and the preliminary observations from initial low load testing. The critical technical challenge, namely the rotating liner face seal, appears to be operating with negligible gas leakage and within the hydrodynamic lubrication regime for the loads tested (peak cylinder pressures of the order of 100 bar) and up to about 10 bar BMEP (brake mean effective pressure). Preliminary testing has proven that the metal-to-metal contact in the piston assembly mostly vanished, and a friction reduction at idle conditions of about 40% as extrapolated to a complete engine has taken place. It is expected that as the speed increases, the friction reduction percentage will diminish, but as the load increases, the friction reduction will increase. The fuel economy benefit over the US Heavy-Duty driving cycle will likely be of the order of 10% compared to a standard engine. Citation: Dardalis, D.; Basu, A.; Hall, M.J.; Mattthews, R.D. The Rotating Keywords: engine efficiency; engine friction; engine lubrication; carbon emissions reduction Liner Engine (RLE) Diesel Prototype: Reducing Internal Engine Friction by about 40% under Idle Conditions. Appl. Sci. 2021, 11, 779. 1. Introduction https://doi.org/10.3390/app11020779 The Rotating Liner Engine (RLE, www.rotatingliner.com) is a unique lubrication

Received: 30 November 2020 concept aimed at Diesel engines, where the cylinder liner rotates to eliminate the boundary Accepted: 12 January 2021 friction environment for the piston rings and skirt around the TDC (top dead center) of Published: 15 January 2021 compression-expansion. It is well understood from the literature that this speciﬁc form of friction loss is Publisher’s Note: MDPI stays neu- responsible for a substantial portion of the mechanical losses in Diesel engines (especially tral with regard to jurisdictional clai- at high loads). In fact, the observed considerable increase in FMEP (friction mean effective ms in published maps and institutio- pressure) as a function of engine load can be traced mostly to this particular factor and it is nal afﬁliations. also responsible for the principal wear mechanism that limits engine life.

1.1. Methods for Measuring Instantaneous Piston Assembly Friction Part of the challenge of evaluating the contribution of the piston ring and skirt bound- Copyright: © 2021 by the authors. Li- censee MDPI, Basel, Switzerland. ary friction is the fact that it is almost impossible to measure in a practical firing engine. The This article is an open access article floating liner method is one of the methods that has shown promise in isolating this term, distributed under the terms and con- but unfortunately, the increasing pressures also affect the sealing mechanism of the floating ditions of the Creative Commons At- liner while piston side loads also interfere with the axial force balance via deflections and tribution (CC BY) license (https:// friction. As a result, the measured piston assembly liner force does not necessarily represent creativecommons.org/licenses/by/ the actual friction. Another factor that limits the accuracy of floating liner testing is the fact 4.0/). that the state of wear of the cylinder liners and piston rings is generally in the close to new

Appl. Sci. 2021, 11, 779. https://doi.org/10.3390/app11020779 https://www.mdpi.com/journal/applsci Appl. Sci. 2021, 11, x FOR PEER REVIEW 2 of 32

necessarily represent the actual friction. Another factor that limits the accuracy of floating Appl. Sci. 2021, 11, 779 2 of 31 liner testing is the fact that the state of wear of the cylinder liners and piston rings is generally in the close to new condition with the honing marks still fresh, and the piston ring profile is still mostly intact as it was optimized by its designers. However, a real heavy- conditionduty engine with as the it operates honing marks in the still field fresh, will and reach the a piston state relatively ring profile early is still in mostlyits useful intact life aswhere it was the optimized honing marks by its designers.have disappeared However, from a real much heavy-duty of the top engine pistonas ring it operates stroke, and in thethe field piston will rings reach have a state worn relatively out and reached early in thei its usefulr steady-state life where profile. the honing It would marks be imprac- have disappearedtical to run the from floating much liner of the for top sufficient piston ring hours stroke, (possibly and theover piston 1000) ringsin order have to wornreachout this andstate. reached Therefore, their this steady-state method has profile. value It in would isolating be impractical changes in tothe run frictional the floating characteristics liner for sufficientwith different hours operating (possibly overor design 1000) inparamete order tors. reach However, this state. these Therefore, results may this methodnot be suffi- has valueciently in accurate isolating to changes draw conclusions in the frictional on the characteristics absolute energy with differentcontribution operating of different or design fric- parameters.tion terms. However, these results may not be sufficiently accurate to draw conclusions on theAnother absolute method energy that contribution has been ofproposed different in friction the past terms. to isolate the individual compo- nentsAnother in a firing method engine, that the has P-w been method, proposed was in demonstrated the past to isolate by Marek the individual et al. in 1991. compo- The nentsprimary in a value firing of engine, this approach the P-w is method, that it gives was demonstratedthe overall instantaneous by Marek et friction al. in 1991. torque The of primarya single valuecylinder of this engine approach without is thatany itmodifica gives thetions overall to the instantaneous engine itself. friction One of torque the down- of a singlesides cylinderof this method engine is without that it anyonly modifications worked for low to the load engine cases itself. and Onefor single of the cylinder downsides en- ofgines this methodwith a low is that moment it only workedof inertia. for While low load the cases overall and instantaneous for single cylinder torque engines cannot with by aitself low identify moment and of inertia. separate While individual the overall compon instantaneousents, since we torque know, cannot when by in itselfthe cycle, identify high andinstantaneous separate individual friction torque components, takes place, since we can know, deduce when the in thelikely cycle, source. high Figure instantaneous 1 shows frictiona sample torque of the takes instantaneous place, we friction can deduce torque. the The likely conclusion source. Figurefrom this1 shows work ais sample that the offriction the instantaneous torque is dominated friction torque.by the high-press The conclusionure part from of the this cycle, work where is that the the piston friction as- torquesembly is boundary dominated terms by theare high-pressureactive (note, accord part ofing the to Miura cycle, whereet al [1]., the the piston frictional assembly losses boundaryof the crank terms bearings are active are almost (note, independent according to of Miura instantaneous et al. [1], thebearing frictional load, lossesand are of more the crank bearings are almost independent of instantaneous bearing load, and are more or or less constant throughout the cycle). This trend remains even when the engine load is less constant throughout the cycle). This trend remains even when the engine load is zero, zero, but not as pronounced as when the engine produces torque. This was the case, de- but not as pronounced as when the engine produces torque. This was the case, despite spite the fact that the oil sump temperatures for those tests were fairly low, amplifying the fact that the oil sump temperatures for those tests were fairly low, amplifying the the hydrodynamic terms. At increasing loads, the dominating spike increases, increasing hydrodynamic terms. At increasing loads, the dominating spike increases, increasing the the overall FMEP by 20% (despite the rather low brake mean effective pressure (BMEP) overall FMEP by 20% (despite the rather low brake mean effective pressure (BMEP) that that this air-cooled, naturally aspirated Diesel can generate). It can be safely deduced that this air-cooled, naturally aspirated Diesel can generate). It can be safely deduced that the the piston assembly boundary friction (plus the injector pump, but in these type of en- piston assembly boundary friction (plus the injector pump, but in these type of engines, the gines, the injection pressure was fairly low and, therefore, this contribution was small) is injection pressure was fairly low and, therefore, this contribution was small) is the primary the primary contributor for this 20% increase in total FMEP. It can also be deduced that contributor for this 20% increase in total FMEP. It can also be deduced that while this termwhile is this important term is even important at idle, even at the at muchidle, at higher the much BMEP higher that modern BMEP that heavy-duty modern Dieselheavy- enginesduty Diesel operate, engines this operate, term has this a considerable term has a considerable contribution contribution and possibly and even possibly dominates even thedominates mechanical the losses.mechanical losses.

Figure 1. Instantaneous frictional torque of a single cylinder Diesel [2]. This particular case examined the effect of lubricant viscosity in instantaneous friction. The unclear legends in the ﬁgure refer to the oil temperature (57 ◦C) and the load.

An additional method of measuring friction losses in ﬁring Diesels is the measurement of IMEP (indicated mean effective pressure) and BMEP. This is also not an easy Appl. Sci. 2021, 11, x FOR PEER REVIEW 3 of 32

Figure 1. Instantaneous frictional torque of a single cylinder Diesel [2]. This particular case examined the effect of lubricant viscosity in instantaneous friction. The unclear legends in the figure refer to the oil temperature (57 °C) and the load. Appl. Sci. 2021, 11, 779 3 of 31 An additional method of measuring friction losses in firing Diesels is the measurement of IMEP (indicated mean effective pressure) and BMEP. This is also not an easy measurement since, especially at high loads, the FMEP is the difference between two large measurement since, especially at high loads, the FMEP is the difference between two numbers, and small errors in IMEP can cause large errors in FMEP. Additionally, modern large numbers, and small errors in IMEP can cause large errors in FMEP. Additionally, enginesmodern operate engines under operate very underhigh pressures, very high and pressures, the resulting and the deformation resulting deformation of the parts of the underparts high under loads high can loadsmove canthe movelocation the of location TDC, which of TDC, is critical which to is criticalthe accuracy to the of accuracy the IMEPof measurements. the IMEP measurements. Nevertheless, Nevertheless, the empirica thel model empirical originally model proposed originally by proposed Chen et by al. (1965)Chen [3] et al.is extensively (1965) [3] is used extensively in the indust usedry in and the is industry calibrated and by isusing calibrated the above by IMEP- using the FMEPabove method. IMEP-FMEP In the calibrations method. In presented the calibrations by Dardalis presented et al. 2012 by Dardalis (given to et us al. by 2012 South- (given westto Research us by Southwest Institute), Research the piston Institute), assembly the boundary piston assembly terms can boundary account terms for at canleast account 25 kPa forof FMEP at least at 25 idle kPa and of FMEP 70 kPa at at idle high and loads. 70 kPa This at highseems loads. to be This consistent seems toin bemagnitude consistent in withmagnitude the conclusions with thethat conclusions one can draw that from one canthe work draw by from Marek the work et al. by Marek et al. If theIf RLE the RLE concept concept indeed indeed eliminates eliminates the bo theundary boundary term term from from the piston the piston assembly, assembly, it it maymay shed shed some some additional additional light light onto onto the thecontribution contribution of the of theboundary boundary piston piston assembly assembly termterm in a infiring a ﬁring diesel diesel engine, engine, since since we will we willbe able be able to compare to compare the baseline the baseline and andRLE RLE under under firingﬁring conditions. conditions.

1.2. 1.2.Mechanism Mechanism for Eliminating for Eliminating the Boundary the Boundary Piston Piston Assembly Assembly Term Term by Rotating by Rotating the Liner the Liner The Thenew new lubrication lubrication mechanism mechanism for forthe the RLE RLE is generated is generated by by the the continuous continuous motion motion of of thethe cylinder cylinder liner liner relative relative to to the the piston piston pl plusus the the non-parallel non-parallel mi micro-scratchescro-scratches generated generated by by thethe relative relative orbital orbital motion motion (Figure (Figure 2),2), as described by Dardalis et al. 2012.

Figure 2. Comparison of the micro-scratch structure in the Rotating Liner Engine (RLE) and a conventional engine on the cylinder liner close to the piston ring reversal.

An additional mechanism is proposed here. It is generally recognized that the frictional losses in engines are directly correlated to the high temperature and high shear rate of the lubricant. As the hot oil is sheared by the fast sliding and low ﬁlm thickness surfaces, it undergoes a temporary reduction in operating viscosity, and it is this reduced viscosity Appl. Sci. 2021, 11, 779 4 of 31

that determines the lubrication conditions of the components, both film thickness and friction coefficient. Under high shear rate, the molecules of the lubricant align with the predominant direction of sliding, reducing the actual viscosity. This reduction can be beneficial in areas where pure hydrodynamic lubrication is taking place. However, it can also be detrimental in the areas close to transition from hydrodynamic into mixed. Given that the Stribeck diagram determines a potential increase in the friction coefficient between boundary and hydrodynamic by over two orders of magnitude, it can be deduced that the above factor can be critical in the areas close to TDC when the piston is moving slowly, yet the cylinder pressure is very high. This mechanism may be considerably different, close to the piston reversal area of a rotating liner engine, as the rotating liner allows a continuous change in shear rate direction of the lubricant on the cylinder walls, and therefore no complete alignment of the lubricant molecules to the predominant shear rate direction. That is, as the piston is ascending, the molecules of the oil on the liner are forced to align in a certain direction offset from the vertical direction as in a conventional engine. However, as the piston is descending, the predominant shear rate is now at an angle to the one already laid on the cylinder wall, forcing an increase in the effective viscosity, which will be of great benefit in the mixed regime area (in the hydrodynamic zone, since the linear piston speed dominates, the increase in viscosity is likely low). This mechanism likely contributes to retaining the hydrodynamic film in the TDC area, both for the RLE, but also for the SVE (Sleeve Valve Engine). The above theory can be extended and slightly modified for the piston ring lubrication very near or right at TDC when the dominating motion is very close to pure rotation. In addition to the above mechanisms (micro-scratch, shear rate directional changes, squeeze film normal to predominant shear rate), we have another factor. The very high gas pressure is also trying to generate a shear rate (in axial direction) which is also perpendicular to the predominant shear rate (note that in the conventional engine, all these directions are perfectly aligned, ensuring that the oil molecules also remain aligned). Note that this factor is far more important in the barrel face ring as opposed to the face seal because the effective sealing zone is actually very narrow, generating a high pressure gradient (dp/dy) in the area of low film thickness. It can be expected that this “conflict” in shear rate direction will also amplify the viscosity and help minimize or eliminate contact in the areas very close to TDC where liner sliding speed dominates over piston axial motion. This can be important, because if significant contact took place in that area, the instantaneous liner driving torque requirement would increase considerably, and FMEP wear would also increase (both events are proven not to be the case from Dardalis et al. 2005 [4], where liner torque remained very low at peak pressure, and of course, the SVE experience where wear rate near TDC was almost eliminated). This will also affect RLE face seal lubrication, but in a slightly different way (the reader is encouraged to review the face seal geometry as shown by Dardalis et al. 2012 [5]). The analyses shown by Dardalis et al. (2012) show that squeeze film lubrication plays a considerable role in the primary sealing ring operating film thickness. The squeeze film is generating a shear rate perpendicular to the predominant shear rate due to rotation, which can also be responsible for localized viscosity increase. This phenomenon would also take place in journal bearings close to the edges; however, in journal bearings, much of the predominant squeeze film direction is also parallel to the sliding direction. In this face seal geometry, the two directions are perfectly perpendicular, since there is no variation in film thickness along the perimeter. Therefore, the pressure tolerance of the RLE face seal may prove higher than predicted by Dardalis et al. (2012).

1.3. Lubricant Starvation for Top Ring Lubrication It has been suggested to the authors that one of the factors that generate boundary lubrication in the top ring TDC reversal is the oil starvation in the top part of the liner. However, the amount of lubricant necessary to achieve the minimization of metallic contact Appl. Sci. 2021, 11, 779 5 of 31

is very small. The authors have reasons to believe that the observed starvation at TDC may be more the result of metallic contact and resulting oil scraping rather than the cause of metallic contact (for example, we know that there is some boundary friction at BDC when there is no starvation issue). Lubricant starvation in the TDC area would in fact generate more friction and wear in engines with rotating liners. Based on the following section, and based on the preliminary testing of this current prototype, this does not seem to be the case.

1.4. Prior Engines with Moving Liners and Prior RLE Results A detailed description of prior engines with partially rotating cylinders, Sleeve Valve Engines, has been provided by Dardalis et al. (2012). A lengthy review will not be presented here. However, the reader needs to know that the aircraft SVEs operated under BMEPs comparable to heavy-duty diesels, but the cylinder rotation during the high-pressure part of the cycle eliminated the typical wear pattern in the TDC area and also reduced overall Appl. Sci. 2021, 11, x FOR PEER REVIEW 6 of 32 FMEP, despite the fact that the sleeve valve added considerably more rotating components than the conventional valve train that it replaced. The reason for this decrease in friction while adding more sources of friction was initially mystifying. Ricardo (1968) eventually assemblydetermined friction that appears the decrease lower in than FMEP its true was contribution. due to the motion A good of theexample “sleeve” of this (liner) is pre- near sentedthe TDC, by Stanton which minimized (2013). In thethe presenceStanton case, of metal-to-metal removal of the contact cylinder from head the piston generated assembly a reductioncomponents in FMEP around of the TDCorder of of compression–expansion. 8 psi (55 kPa), while the final removal of the pistons generatedThe a ﬁrst reduction RLE prototype of 2–5 psi was (14–34 a single kPa) cylinderin FMEP. conversion (Note, in the of Stanton a GM Quad (2013) 4 light-dutytest, the valveSI engine train was [4]. Unlikeentirely this removed current by Diesel removing prototype, the cylinder that light-duty head, but conversion disabling was the designed valve trainsuch while that the the head space remains of the neighboringin place has an cylinders identical was effect.) invaded In the by typical the RLE published components, tear downcondemning test, this procedure the engine is into natural a single because cylinder. it is difficult The face to run seal, the while engine similar removing in overall the pistonfunction, and connecting had certain rod design while limitations the valve which train is were still correctedactive. It forwould the currentrequire heavy-dutythe use of bobdesign. weights, However, which thewere initial present design in allowedthe work hot presented motoring by tear Dardalis down tests,et al [5] which and facilitated Kim et al the[6]. measurementIt needs to be noted of friction that the characteristics. valve train friction The rather is not large affected motoring by the FMEPpresence reduction, of the pistonalong (as with long a tearas the down oil temperature test to conﬁrm is maintained that the difference at the same was fromlevel theof course), piston assembly,while the is pistonshown assembly in Figure friction3. will decrease dramatically in the absence of a working valve train.

FigureFigure 3. Hot 3. Hotmotoring motoring friction friction of the of light-duty the light-duty RLE compared RLE compared to baseline, to baseline, borrowed borrowed from Kim from et al.Kim (2005). et al.The (2005). x axisThe shows x axis the showsengine thespeed, engine and speed,the y axis and shows the y the axis measured shows the motoring measured FMEP motor- foring the FMEP two engines. for the two This engines. was performed This was after performed the tear after down the process tear down allowe processd the allowed isolation the of isolationthe piston assembly FMEP from the other terms (i.e., crankshaft, camshaft, etc.). of the piston assembly FMEP from the other terms (i.e., crankshaft, camshaft, etc.).

1.5. ReducedThe authorsLubricant have Viscosities received for Heavy-Duty questions about Engines the relatively large amount of FMEP reductionOne modern presented trend in of Figure reducing3, especially friction inin lightheavy-duty of tear downDiesels, tests especially performed those by otherthat areresearchers subjected to that fuel show efficiency the overall standards, FMEP is contribution the reduction of in the lubricant piston assembly viscosity. to This be lowerap- proachthan theis well reduction captured shown in the in results Figure 3by. TheMare precisionk et al. (1991) of these [2]. tests Their by results Kim et clearly al. (2005) show [ 6 ] thatis notreduced likely lubricant to be the viscosity issue, since reduces these the measurements friction in the wereintake extremely and exhaust straight strokes, forward but has no effect in the late compression/early expansion, even though the higher viscosity may offer some relief in the metal-to-metal contact intensity for some areas for some portion of the cycle (and therefore, reduced wear compared to the lower viscosity). This is consistent with many recent publications on heavy-duty lubricant formulations with low viscosity lubricants, where measurable improvement is shown for low load operating conditions, but only modest or negligible improvement at high loads. The reason cited is sometimes the misconception about the presence or the significance of the mixed lubrication in the high-pressure part of the cycle, but the real reason seems to be the fact that the lubricant viscosity plays a limited role in the actual fMEP during the loaded part of the cycle because the lubrication regime is deep into the boundary area where the friction coefficient is flat in the Stribeck diagram. However, viscosity always seems to play a leading role during the exhaust and intake strokes.

Appl. Sci. 2021, 11, 779 6 of 31

and conducted with two independent measuring techniques, while the margin of error is considerably smaller than the FMEP improvement shown. The authors are convinced that the primary source of the misconception is that most of the published tear down tests (particularly on light-duty engines such as the one tested by Kim et al. [6]) are almost always performed in the wrong order. The valve train is disabled before the piston components are removed, and this difference in torque is assigned entirely to the valve train. This is an erroneous conclusion, because when the valve train is disabled, the piston ring gas pressure loading is also eliminated, and therefore, piston assembly boundary friction is mostly eliminated. The result of this is that the observed reduction in friction by removal or disabling of the valve train, while caused by the fact that the piston assembly is no longer loaded by gas pressure, is erroneously assigned to the valve train. Therefore, piston assembly friction appears lower than its true contribution. A good example of this is presented by Stanton (2013). In the Stanton case, removal of the cylinder head generated a reduction in FMEP of the order of 8 psi (55 kPa), while the ﬁnal removal of the pistons generated a reduction of 2–5 psi (14–34 kPa) in FMEP. (Note, in the Stanton (2013) test, the valve train was entirely removed by removing the cylinder head, but disabling the valve train while the head remains in place has an identical effect.) In the typical published tear down test, this procedure is natural because it is difﬁcult to run the engine removing the piston and connecting rod while the valve train is still active. It would require the use of bob weights, which were present in the work presented by Dardalis et al. [5] and Kim et al. [6]. It needs to be noted that the valve train friction is not affected by the presence of the piston (as long as the oil temperature is maintained at the same level of course), while the piston assembly friction will decrease dramatically in the absence of a working valve train.

1.5. Reduced Lubricant Viscosities for Heavy-Duty Engines One modern trend of reducing friction in heavy-duty Diesels, especially those that are subjected to fuel efficiency standards, is the reduction in lubricant viscosity. This approach is well captured in the results by Marek et al. (1991) [2]. Their results clearly show that reduced lubricant viscosity reduces the friction in the intake and exhaust strokes, but has no effect in the late compression/early expansion, even though the higher viscosity may offer some relief in the metal-to-metal contact intensity for some areas for some portion of the cycle (and therefore, reduced wear compared to the lower viscosity). This is consistent with many recent publications on heavy-duty lubricant formulations with low viscosity lubricants, where measurable improvement is shown for low load operating conditions, but only modest or negligible improvement at high loads. The reason cited is sometimes the misconception about the presence or the significance of the mixed lubrication in the high-pressure part of the cycle, but the real reason seems to be the fact that the lubricant viscosity plays a limited role in the actual fMEP during the loaded part of the cycle because the lubrication regime is deep into the boundary area where the friction coefficient is flat in the Stribeck diagram. However, viscosity always seems to play a leading role during the exhaust and intake strokes.

1.6. Lubricant Formulation in RLE Commercial Units In addition to generating experimental proof of the potential contribution of the boundary and mixed lubrication in diesel engines, the authors also expect that, given the very high predicted reduction in FMEP compared to conventional methods, and given the relatively moderate increase in engine cost, the RLE will ﬁnd its way in Heavy-Duty commercial engines. If that is to take place, the question that arises is what type of lubricant will be ideal for the RLE? Would the standard 15w40 lubricant be ideal, or should RLE units follow the trend of reduced viscosity? For the testing of the ﬁrst light-duty RLE prototype (Kim et al. 2005) [6], the standard 5w30 low-viscosity light-duty lubricant was used. The reason for choosing the above grade was that it was expected that the low viscosity would amplify the boundary terms for the Appl. Sci. 2021, 11, 779 7 of 31

low-speed un-throttled motoring conditions. Additionally, since the peak pressures for motoring were relatively low (just under 20 bar), the low viscosity oil would not provide any challenges for lubricating the RLE face seal. During the initial testing of the RLE diesel prototype, the 15w40 grade has been consistently used. Dardalis et al. (2012) [5] presented calculations that showed that if a high-temperature, high-shear rate lubricant of about 0.0035 Pa*s is used (lower than the one exhibited by 15w40 and about identical to the modern low-viscosity oils currently proposed for Heavy-Duty use), then the seal film thickness will be adequate to avoid contact. Even though the low load cases that we have been running so far in the prototype would not exceed the limit of the seal calculated by Dardalis (2012) even with a 5w30 oil, and even though we expect that these calculations were conservative for the above reasons and computational complications discussed by Dardalis et al. (2012), the security that the high-viscosity oil provides is considered critical for the prototype’s preservation. Of course, once more data are accumulated, and the prototype is no longer as precious to us as it is now, lower viscosity can be tried as well so that more can be learned about RLE operation under reduced viscosity. When the conversion is redesigned for larger engines, the bearing surfaces of the RLE face seal can be redesigned larger to accommodate for the lower viscosity with a higher confidence level. However, it is not clear yet if the lower viscosity may require higher liner speeds to achieve the same results in minimizing the boundary terms of the piston rings, especially when operating under the high pressures of the future engines. If that proves to be the case, the 15w40 grade may prove more desirable. Furthermore, most of the benefits of the low-viscosity oil can be gained by the 15w40 grade, if one simply increases the clearance in the crankshaft bearings and piston skirts. This would allow exploitation of the high viscosity for the top-end cylinder and RLE seal lubrication, while the rest of the hydrodynamic terms can still benefit from the reduced shear rate. It needs to be noted that in any hydrodynamic bearing design, the compromise between reduced viscous losses and reduced maximum load capacity work in an identical manner regardless whether the designer reduces the viscosity or increases clearance. Due to the recent regulations in fuel economy for Heavy-Duty truck engines in North America, a new trend of reduced lubricant viscosities has emerged. As discussed above, and supported by Marek et al., the reduced viscosity will reduce the friction in the intake and exhaust stroke, while having almost no effect in the high-pressure part of the cycle, which seems to dominate the mechanical losses. Therefore, as expected, the benefit is marginal in terms of fuel economy, while introducing a risk of potential increased wear. In order to control the wear impact on the piston rings and cylinder liner, more aggressive and expensive additive packages have been developed. Regardless of what grade viscosity will be used in the RLE when it reaches production, given that the metallic intensity will be greatly minimized (which may also be accompanied by a significant reduction in piston ring temperature), many of these additives can be reduced or eliminated. This could not only reduce the cost of the lubricant, but also increase the life of the catalysts, which are known to degrade faster in the presence of large amounts of lubricant additives.

2. The Heavy-Duty RLE Prototype 2.1. Selected Pictures from the Diesel RLE Prototype This section presents pictures of the prototype. The actual CAD of the prototype was presented by Dardalis et al. 2012 [5]. Relatively minor design changes have been made, and many of these will be pointed out in this section. Both baseline and RLE started life as a four-cylinder Cummins 4BT 3.9 model year prior to 1998 (two valves per cylinder, mechanical fuel injection). Both are converted to single-cylinder operation. The overall layout of the RLE prototype can be seen in Figure 4. Cylinder #2 is the only active cylinder, and all other three crank throws are covered by bob weights that Appl. Sci. 2021, 11, 779 block oil leakage and also rebalance the crankshaft. The brass driving gear on8 ofthe 31 position of cylinder #1 engages the crankshaft via an external bevel 90-degree gearbox and belt with a total gear reduction of about 2.7:1 (later, when detailed testing was performed, the ratio was reduced to 3:1). This driving gear is identical in diameter and number of teeth total gear reduction of about 2.7:1 (later, when detailed testing was performed, the ratio wasto the reduced rotating to 3:1).liner This flange driving gear; gear therefore, is identical it could in diameter be replaced and numberby an actual of teeth rotating to liner theif the rotating engine liner was ﬂange to be gear; a complete therefore, four-cylinder it could be replaced (proposed by an driving actual rotating mechanism liner if is shown theby engineDardalis was et to al. be 2012). a complete However, four-cylinder now two (proposed bearing drivingholders mechanism are secured is shownon the byuntouched Dardalisoriginal etengine al. 2012). bore, However, and these now bearings two bearing support holders the arevertical secured shaft on that the untouchedis shown in Figure original4. The engineshaft goes bore, andthrough these bearingsthe cylinder support head the verticaland engages shaft that the is shownbevel ingearbox Figure4 discussed. Theabove. shaft goes through the cylinder head and engages the bevel gearbox discussed above.

FigureFigure 4. 4.Block Block of of the the RLE RLE diesel diesel prototype. prototype. The driving The drivin gearg could gear be could replaced be replaced by another by rotating another rotat- liner,ing liner, so the so engine the engine could operate could operate as a complete as a complete four-cylinder four-cylinder 4BT RLE. 4BT RLE.

FigureFigure5 shows 5 shows the the details details of the of face the sealface features seal features on the rotatingon the rotating liner flange. liner These flange. These areare exactly exactly as as described described by Dardalisby Dardalis et al. et 2012. al. 201 The2. face The has face been has flattened been flattened by lapping by tolapping to thethe tolerance tolerance level level of typicalof typical face face seals, seals, which which is of theis of order the oforder 1 micron of 1 micron flatness flatness or less, or less, and was also polished. There are nine oil supply holes across the flange which feed oil and was also polished. There are nine oil supply holes across the flange which feed oil to to the annular groove. The annular area inside the annular groove is called the “sealing zone”the annular (see Dardalis groove. 2012) The [5 ]annular and acts area as a sealinginside damthe annular and squeeze groove film is bearing. called Thethe “sealing etchedzone” step (see pads Dardalis are visible, 2012) and [5] these and constitute acts as a the sealing primary dam hydrodynamic and squeeze features. film bearing. The The primaryetched step sealing pads ring are can visible, also be and seen, these held cons by atitute human the hand. primary The hydrodynamic primary sealing ringfeatures. The isprimary actually sealing flat (the ring radial can marks also seenbe seen, are fromheld theby oil,a human which hand. is deposited The pr inimary an uneven sealing ring is manner,actually replicating flat (the radial the pattern marks from seen the are rotating from linerthe oil, features which that is itdeposited was in contact in an with). uneven man- Thener, head replicating gasket is the a 2.25 pattern mm (0.090from the inch)-thick rotating copper liner features unit (identical that it gasket was in used contact in the with). The Appl. Sci. 2021, 11, x FOR PEER REVIEWbaseline). In the case of the RLE, the head gasket only seals coolant and oil, so the head 9 of 32 head gasket is a 2.25 mm (0.090 inch)-thick copper unit (identical gasket used in the base- bolts need not be torqued down to the full 90 ft-lbs required by the Cummins shop manual. line). In the case of the RLE, the head gasket only seals coolant and oil, so the head bolts need not be torqued down to the full 90 ft-lbs required by the Cummins shop manual.

FigureFigure 5. Details5. Details of the of rotatingthe rotating liner faceliner seal face features. seal featur The copperes. The head copper gasket head only gasket needs only to seal needs oil to seal andoil coolant.and coolant. This picture This picture is shown is againshown in theagain Appendix in the AppendixA with a higher A with resolution. a higher resolution.

Figure 6 shows the block with the cylinder head removed, the rotating liner pulled out to expose the polished outside surface. The brass thrust washer on the head block is visible. This thrust washer is loaded by the spring load acting on the primary sealing ring. However, as discussed by Dardalis et al. (2012) [5], when the engine is started and oil pressure is restored, the oil pressure will take the load and this thrust washer will no longer be active. The bearing holder for the driving gear shaft but without the ball bearing used is also visible on the bottom left of the picture (the driving gear has been removed). Figure 7 shows the cylinder head with the head insert and primary sealing ring springs (component terminology follows Dardalis et al. 2012). This part is built out of mild steel and houses the secondary seal and springs, as described by Dardalis et al. 2012 [5]. The only deviation from the Dardalis et al. (2012) description is that conventional coil springs are used in place of a conical washer spring due to cost considerations. We are also using two flywheels instead of one in order to minimize the crank speed variations typical in single-cylinder engines and to make engine starting easier.

Figure 6. Rotating liner pulled a few inches out of the block, exposing the outside surface.

Figure 7. Cylinder head, showing the head insert that houses the primary sealing ring, springs that pre-load the primary sealing ring, and secondary seal. Appl. Sci. 2021, 11, x FOR PEER REVIEW 9 of 32

Appl. Sci. 2021, 11, x FOR PEER REVIEW 9 of 32

Figure 5. Details of the rotating liner face seal features. The copper head gasket only needs to seal oil and coolant. This picture is shown again in the Appendix A with a higher resolution.

Figure 6 shows the block with the cylinder head removed, the rotating liner pulled out to expose the polished outside surface. The brass thrust washer on the head block is

visible. This thrust washer is loaded by the spring load acting on the primary sealing ring. FigureHowever, 5. Details as discussed of the rotating by Dardal liner faceis etseal al. featur (2012)es. [5],The copperwhen thehead engine gasket onlyis started needs andto seal oil oil and coolant. This picture is shown again in the Appendix A with a higher resolution. Appl. Sci. 2021, 11, 779 pressure is restored, the oil pressure will take the load and this thrust washer9 of 31 will no longer be active. The bearing holder for the driving gear shaft but without the ball bearing Figure 6 shows the block with the cylinder head removed, the rotating liner pulled used is also visible on the bottom left of the picture (the driving gear has been removed). out to expose the polished outside surface. The brass thrust washer on the head block is Figure 7 shows the cylinder head with the head insert and primary sealing ring visible.Figure This6 shows thrust the washer block withis loaded the cylinder by the headspring removed, load acting the rotating on the primary liner pulled sealing ring. springs (component terminology follows Dardalis et al. 2012). This part is built out of mild outHowever, to expose as the discussed polished by outside Dardal surface.is et al. The (2012) brass thrust[5], when washer theon engine the head is started block and oil steel and houses the secondary seal and springs, as described by Dardalis et al. 2012 [5]. ispressure visible. This is restored, thrust washer the oil is loadedpressure by thewill spring take loadthe load acting and on thethis primary thrust sealingwasher will no The only deviation from the Dardalis et al. (2012) description is that conventional coil ring.longer However, be active. as discussed The bearing by Dardalis holder for et al.the (2012) driving [5], gear when shaft the engine but without is started the and ball bearing oilsprings pressure are is used restored, in place the oil of pressure a conical will wa takesher the spring load anddue thisto cost thrust considerations. washer will no used is also visible on the bottom left of the picture (the driving gear has been removed). longerWe be active. are also The using bearing two holder flywheels for the instead driving gearof one shaft in order but without to minimize the ball bearingthe crank speed Figure 7 shows the cylinder head with the head insert and primary sealing ring usedvariations is also visibletypical on in the single-cylinder bottom left of theengines picture and (the to driving make gearengine has starting been removed). easier. springs (component terminology follows Dardalis et al. 2012). This part is built out of mild steel and houses the secondary seal and springs, as described by Dardalis et al. 2012 [5]. The only deviation from the Dardalis et al. (2012) description is that conventional coil springs are used in place of a conical washer spring due to cost considerations. We are also using two flywheels instead of one in order to minimize the crank speed variations typical in single-cylinder engines and to make engine starting easier.

FigureFigure 6. 6.Rotating Rotating liner liner pulled pulled a few a inchesfew inches out of out the of block, the exposingblock, exposing the outside thesurface. outside surface.

Figure7 shows the cylinder head with the head insert and primary sealing ring springs (component terminology follows Dardalis et al. 2012). This part is built out of mild steel and houses the secondary seal and springs, as described by Dardalis et al. 2012 [5]. The only deviation from the Dardalis et al. (2012) description is that conventional coil springs areFigure used 6. in Rotating place of liner a conical pulled washer a few spring inches dueout toof costthe block, considerations. exposing the outside surface.

Figure 7. Cylinder head, showing the head insert that houses the primary sealing ring, springs that pre-load the primary sealing ring, and secondary seal.

FigureFigure 7. 7.Cylinder Cylinder head, head, showing showing the head the head insert insert that houses that houses the primary the primary sealing ring, sealing springs ring, that springs pre-loadthat pre-load the primary the primary sealing ring, sealing and ring, secondary and secondary seal. seal.

We are also using two ﬂywheels instead of one in order to minimize the crank speed variations typical in single-cylinder engines and to make engine starting easier.

2.2. Fuel Injection System The injection system in use is the old mechanical type, mainly due to our very low budget. The same injection pump and injector are used for both engines. Partly due to the conversion of the injection pump operating only one cylinder, and partly because single cylinder engines are difﬁcult to start, the injection timing is set fairly advanced. Additionally, the cracking pressure of the injector used is set about 30% below the original manufacturer settings. This applies to both baseline and RLE. This generates very high peak cylinder pressures during idle conditions. When the engines are cold, peak pressures are Appl. Sci. 2021, 11, 779 10 of 31

of the order of 67–70 bar. While this is probably very inefficient from the thermodynamic standpoint, this would be irrelevant to this investigation as long as both engines operate with the same injection timing. One added benefit of this factor is that we have the opportunity to test the RLE seal at pressure conditions more severe than the load conditions currently used. A second benefit is that the fueling rate may have a much larger impact on peak cylinder pressure because much of the fuel is consumed near the TDC at the low loads that we are currently working with, and since peak cylinder pressure is far easier to measure than fueling rate, some preliminary assessment of the fueling rate between the two engines may be readily derived. This has been expanded upon more in the next section. However, there is an overwhelming downside to the old technology injection system. That is, we have very limited control over the combustion and as a result, we have to follow a “hit and miss” approach in generating comparable conditions, as again illustrated on following sections.

2.3. Face Seal Design Challenge: Experimental Confirmation As discussed by Dardalis et al. (2012) [5], the RLE seal needs to have negligible gas leakage, negligible wear, and very low added friction. Dardalis et al. (2012) described in detail the design and the philosophy behind the RLE face seal. A relatively low closing force compared to the bearing area is used so that very high cylinder pressures can be contained without metal-to-metal contact. Significant oil flow is established throughout the primary sealing ring, in order to keep it cool and also in order to ensure that sufficient oil is present to block the gas flow between the gap (film thickness ranges from up to 5 microns in the unloaded part of the cycle to under 1 micron under full load peak pressure). The abundance of lubricant also allows for a very low surface roughness to be used, which allows a much lower minimum film thickness prior to metal-to-metal contact compared to the piston rings that ride on a relatively high surface roughness. However, the fairly high spring pre-load and very tight flatness tolerance also ensure that oil leakage into the combustion chamber is negligible. As described by Dardalis et al. (2012) [5], negligible means that in the worst case scenario, it would be a tiny fraction of the lubricant deposited on the liner by the piston rings during each stroke. However, based on the calculation assumptions presented by Dardalis (2012), there may be almost zero net flow, given that the gas pressure will push much of this oil back from the sealing zone. Based on the preliminary testing, it appears that these requirements have been met. The RLE and BSL (Baseline) seem to have an approximately identical starting performance. There is no indication of blowby from the RLE from exhaust port #1, which is directly open to the atmosphere and isolated from the remaining crank case. However, the best indication of zero blowby from the RLE face seal is the trace of carbon. Given the poor fuel atomization of our fuel injection system, a significant amount of carbon is deposited on all surfaces that are in contact with combustion products within a few minutes of engine operation. However, the surfaces of the rotating liner flange very close to the sealing gap, and head inserts are completely clean from any traces of combustion products. Therefore, no gas flow of any significance has been established while the engine was operating, which is to be expected, given that the RLE face seal has no end-gap. Furthermore, even though we have not measured the liner driving torque for this prototype, it appears to be fairly low, based on the fact that the drive belt (Figure8) is not tight, and yet there is no observed belt slippage (note, this indicates that the pinned piston rings also likely ride hydrodynamically). The relatively soft copper alloy primary sealing ring does not show any signs of wear. This is of course premature, given that the peak pressure so far has not exceeded approximately 100 bar. However, the engine has now operated for about 50 h, has been subjected to about 10 bar BMEP, and we have still seen no malfunction or wear of the sealing system. Appl. Sci. 2021, 11, x FOR PEER REVIEW 11 of 32

there is no observed belt slippage (note, this indicates that the pinned piston rings also likely ride hydrodynamically). The relatively soft copper alloy primary sealing ring does not show any signs of wear. This is of course premature, given that the peak pressure so Appl. Sci. 2021, 11, 779 far has not exceeded approximately 100 bar. However, the engine has now11 operated of 31 for about 50 h, has been subjected to about 10 bar BMEP, and we have still seen no malfunction or wear of the sealing system.

FigureFigure 8. 8.Overall Overall layout layout of of the the prototype, prototype, seen seen here here running. running. The The 90-degree 90-degree gearbox gearbox is the is total the total transaxletransaxle of of a smalla small tractor, tractor, so itso is it much is much larger larger than than it needs it needs to be. to The be. actual The bevelactual gears bevel used gears in used in therethere are are approximately approximately 3 inches 3 inches in diameter. in diameter. This gearbox This gearbox was later was replaced later byreplaced a smaller by and a smaller more and efﬁcientmore efficient unit. unit.

3.3. Testing Testing the the RLE RLE and and BSL BSL Single Single Cylinder Cylinder Engines Engines under under Idle Conditions Idle Conditions InIn this this section, section, we willwe describewill describe the testing the oftesting the Baseline of the Engine Baseline (identical Engine to the(identical RLE to the but with a conventional cylinder) and the RLE prototype described above. By comparing RLE but with a conventional cylinder) and the RLE prototype described above. By com- the combustion characteristics, we are quantifying the friction reduction experimentally underparing low-speed the combustion and peak characteristics, cylinder pressure we (PCP) are conditions. quantifying This the section friction is rather reduction long experi- formentally the two followingunder low-speed reasons. We and cannot peak accurately cylinder obtain pressure the iMEP (PCP) of theconditions. two engines This and section is werather cannot long control for the the mechanicaltwo following fuel injectionreasons. well. We Acannot measurement accurately of iMEP obtain for the twoiMEP of the enginestwo engines would and have we been cannot a very control direct way the of mechan comparingical thefuel idle injection fMEP, but well. unfortunately, A measurement of thisiMEP is not for feasible. the two Anengines electronically would have controlled been a fuel very injection direct wouldway of have comparing allowed the us toidle fMEP, easilybut unfortunately, compare cases. this Instead, is not a relativelyfeasible. complicatedAn electronically methodology controlled that generatesfuel injection an would evenhave comparison allowed us of to operating easily compare conditions cases. between Instead, the BSL a relatively and RLE iscomplicated presented, which methodology accomplishes the same, but unfortunately is lengthy. that generates an even comparison of operating conditions between the BSL and RLE is Given the fact that both engines are powered by an approximately 1 L single cylinder whichpresented, drives which a 5-bearing accomplishes crankshaft the and sa accessoriesme, but unfortunately sized for a 5.9 is Llengthy. 6-cylinder engine (waterGiven pump the and fact oil that pump both are engines identical are to powere the 6BT),d andby an given approximately that we have 1 L focused single cylinder onwhich 750 rpm, drives the a operating 5-bearing conditions crankshaft presented and accessories here represent sized a loaded for a idle5.9 L (i.e., 6-cylinder engine engine is(water driving pump an AC and compressor oil pump or are other identical accessories) to the rather6BT), thanand given a totally that unloaded we have idle focused on condition750 rpm, of the a fulloperating engine. conditions The reason whypresented we chose here 750 represent rpm is because a loaded the idle minimum (i.e., engine is idledriving of the an two AC engines compressor is on the or orderother ofaccessories) 630 to 680 rpm,rather and than is nota totally entirely unloaded consistent. idle condi- Therefore,tion of a full we areengine. focusing The on reason an idle why condition we ch whereose 750 the rpm fueling is because lever of the the minimum injection idle of pump needs to be slightly advanced instead of resting at its minimum stop. These old-style the two engines is on the order of 630 to 680 rpm, and is not entirely consistent. Therefore, injection pumps operate like speed regulators in the idle range. Additionally, we have focusedwe are onfocusing 70 ◦C because on an theidle original condition BSL datawhere are the at that fueling oil/coolant lever of temperature, the injection and pump we needs haveto be not slightly yet obtained advanced high instead resolution of pressureresting at data its for minimum the BSLat stop. a higher These temperature. old-style injection Additionally,pumps operate all the like data speed are with regulators the 15w40 in gradethe idle lubricant. range. Additionally, we have focused on 70 °C because the original BSL data are at that oil/coolant temperature, and we have not 3.1.yet Preliminaryobtained high Testing resolution from Prior pressure Publication data for the BSL at a higher temperature. Addition- ally,The all fuelthe data economy are with at idle the of 15w40 the prototype grade lubricant. was compared to a baseline engine, also single-cylinder, where all the characteristics were almost identical to the RLE except for the3.1. liner Preliminary rotation (DardalisTesting from 2019). Prior The Publication condition focused on at that time was 750 rpm idle and 71 ◦C coolant and oil temperature. For that condition, the RLE was estimated to have slightlyThe under fuel 20% economy lower fuel at idle consumption. of the prototyp This estimatione was compared was achieved to a via baseline comparing engine, also thesingle-cylinder, pressure rise between where theall the peak characteristics motoring pressure were (PMP) almost and identical peak cylinder to the pressure RLE except for duringthe liner ﬁring rotation pressure (Dardalis (PCP), which2019). isThe essentially condition proportional focused on to at the that total time heat was release 750 rpm idle perand cycle, 71 °C as coolant long as and the combustion oil temperature. phasing For is th similar.at condition, The assumption the RLE herewas isestimated that the to have indicatedslightly under thermal 20% efﬁciency lower offuel the consumption. two engines is This identical, estimation which is was reasonable achieved since via the comparing engines are almost completely identical with similar compression ratio. The validity of this comparison was also made possible by the fact that injection timing was substantially Appl. Sci. 2021, 11, x FOR PEER REVIEW 12 of 32

the pressure rise between the peak motoring pressure (PMP) and peak cylinder pressure during firing pressure (PCP), which is essentially proportional to the total heat release per cycle, as long as the combustion phasing is similar. The assumption here is that the indi- Appl. Sci. 2021, 11, 779 12 of 31 cated thermal efficiency of the two engines is identical, which is reasonable since the engines are almost completely identical with similar compression ratio. The validity of this comparison was also made possible by the fact that injection timing was substantially ad- advancedvanced and and peak peak firing firing pressure pressure was was around around the the TDC, TDC, which which meant meant that the majoritymajority ofof thethe heatheat releaserelease waswas completecomplete byby TDC.TDC. ThisThis isis notnot anan efficientefficient wayway toto runrun anan engine,engine, butbut wewe areare notnot attemptingattempting toto achieveachieve aa recordrecord inin thermalthermal efficiency;efficiency; instead,instead, wewe areare estimatingestimating thethe percentagepercentage ofof frictionfriction reductionreduction ofof thethe RLERLE conceptconcept byby comparingcomparing thethe amountamount ofof heatheat releaserelease neededneeded toto sustainsustain thethe idleidle inin thethe BSLBSL andand RLE.RLE. TheThe differencedifference betweenbetween PCPPCP andand PMPPMP was adjusted adjusted by by the the ratio ratio of of clearance clearance volumes volumes between between the the two two engines engines (which (which un- unfortunatelyfortunately were were not not exactly exactly the the same, same, but but also also varied, varied, as asshown shown in inthe the sections sections below) below) in inorder order to tocomply comply with with the the ideal ideal gas gas equation, equation, and and the the corrected corrected value value is what is usedused forfor thethe finalfinal comparisoncomparison (the(the correctioncorrection factorfactor isis relativelyrelatively small).small). ThisThis yieldedyielded aboutabout aa 19%19% reductionreduction inin fuelfuel consumptionconsumption atat idleidle forfor thethe RLERLE (Dardalis(Dardalis 2019).2019). However,However, thesethese teststests werewere veryvery preliminarypreliminary andand hadhad thethe followingfollowing limitations.limitations. 1.1. DifficultyDifficulty in obtaining the proper TDC location. 2.2. DifficultyDifficulty in obtaining the accurate IMEP. 3.3. Uncertainty in injection timing. 4.4. Baseline had a slightly higher compression ratio. 5.5. The Rotating Liner gearbox was an inefficientinefficient unit withwith greasegrease bushingsbushings ratherrather thanthan roller element bearings, which increased gearbox parasitic losses. FigureFigure9 9 shows shows the pressure vs. crank crank angle angle graphs graphs from from the the two two prototypes prototypes from from the theprior prior publication. publication. In this In thisgraph, graph, the CA the axis CA axisis slightly is slightly shifted shifted in order in order to align to align the pre- the premixedmixed combustion combustion events, events, which which seem seem to be to fair be fairlyly close close anyway. anyway. However, However, it is it fairly is fairly ev- evidentident that that both both cases cases have have a substantially a substantially over over-advanced-advanced timing timing and andthat thatmost most of the of heat the heatrelease release has been has been completed completed by the by TDC. the TDC. In both In both cases, cases, the peak the peak pressure pressure was waswithin within one onedegree degree from from the theTDC. TDC. In Inthe the analyses analyses presented presented in in this this paper, paper, not not only only did we confirmconfirm that,that, butbut wewe alsoalso showedshowed thatthat thethe BSLBSL waswas slightlyslightly moremore retarded,retarded, andand aa fractionfraction ofof itsits heatheat releaserelease hadhad takentaken placeplace afterafter thethe TDC,TDC, changingchanging thethe numbersnumbers somewhat.somewhat.

FigureFigure 9. Pressure vs. vs. crank crank angle angle of of RLE RLE vs. vs. BSL. BSL. Th Thee two two parabolas parabolas in the in the middle middle indicate indicate the the approximateapproximate peakpeak ofof thethe motoringmotoring pressure.pressure.

Firstly,Firstly, wewe reducedreduced thethe compressioncompression ratioratio ofof thethe BSLBSL toto thethe pointpoint wherewhere thethe motoringmotoring pressurepressure waswas slightlyslightly lowerlower thanthan thethe motoringmotoring pressurepressure onon thethe RLE.RLE. EvenEven thoughthough wewe havehave notnot perfectlyperfectly matchedmatched thethe twotwo compressioncompression ratios,ratios, wewe werewere atat leastleast ableable toto bracketbracket thethe RLERLE compressioncompression ratioratio with with the the two two BSL BSL cases. cases. In In the the work work presented presented in this in this paper, paper, we were we were able to locate the TDC with better precision in both engines. However, we were still not able to obtain repeatable IMEP measurements. We identified the root of this problem to be the crankshaft torsional vibrations that take place due to the conversion from a four-cylinder configuration into a single-cylinder engine. The reasonably precise location of the TDC, however, allowed us to better define the location of the premixed combustion with respect to the TDC (which is the effective injection timing). The potential error caused by the Appl. Sci. 2021, 11, x FOR PEER REVIEW 13 of 32

able to locate the TDC with better precision in both engines. However, we were still not able to obtain repeatable IMEP measurements. We identified the root of this problem to be the crankshaft torsional vibrations that take place due to the conversion from a four- Appl. Sci. 2021, 11, 779 cylinder configuration into a single-cylinder engine. The reasonably precise location13 of of the 31 TDC, however, allowed us to better define the location of the premixed combustion with respect to the TDC (which is the effective injection timing). The potential error caused by torsionalthe torsional vibration vibration is negligible is negligible compared compared to the to potential the potential error error in the in iMEP the iMEP calculation. calcula- Thistion. allowed This allowed us to us better to better reﬁne refine our our relative relative gross gross heat heat release release rate rate (PCP (PCP minus minus PMP). PMP). Additionally,Additionally, we we were were able able to calculateto calculate the adiabaticthe adiabatic instantaneous instantaneous heat release heat release rate, which rate, iswhich also relatively is also relatively insensitive insensitive to the exact to the TDC exact location, TDC location, especially especially if the majority if the ofmajority the heat of inputthe heat takes input place takes during place the during premixed the premixed combustion combustion phase (duephase to (due the to low-pressure the low-pressure fuel systemfuel system modiﬁed modified for single-cylinder for single-cylinder operation), operation), and ifand that if eventthat event is substantially is substantially before be- thefore TDC the TDC where where the combustion the combustion chamber chamber volume volume is relatively is relatively large. large. We wereWe were also also able able to to obtain some limited exhaust CO2 concentrations of the two engines. obtain some limited exhaust CO2 concentrations of the two engines.

3.2.3.2. Locating Locating TDC TDC and and IMEP IMEP Measurements Measurements TheThe location location of of the the top top dead dead center center is is critical critical to to the the calculation calculation of of IMEP. IMEP. In In the the prior prior BSLBSL data,data, wewe had attempted attempted to to find find TDC TDC by by motoring motoring the the engine engine and and observing observing the peak the peakpressure pressure location. location. However, However, even eventhat crude that crude method method had failed had failedto produce to produce good results good resultsbecause because the peak the pressure peak pressure location location varied variedover a overrange a rangeof 0.4 degrees of 0.4 degrees (we used (we a used shaft aencoder shaft encoder with a withresolution a resolution of 20 points of 20 per points degree). per degree). In the RLE In theunit, RLE we unit,used a we proximity used a proximitysensor, replacing sensor, replacingthe injector, the and injector, cranked and the cranked engine the while engine collecting while collecting encoder and encoder prox- andimity proximity sensor output. sensor output.The experiment The experiment was repeated was repeated for various for various cranking cranking speeds speeds (as the (aspowering the powering batteries batteries were becoming were becoming depleted), depleted), and the and minimum the minimum proximity proximity sensor sensor output outputwas repeatable was repeatable at a specific at a specific offset from offset the from Z-pulse the Z-pulse of our ofencoder. our encoder. WhileWhile the the above above method method worked worked well, well, we we still still had had serious serious torsional torsional vibration vibration prob- prob- lemslems (aggravated (aggravated by by the the liner liner mechanism mechanism for for the the RLE, RLE, Figure Figure 10 10)) that that prevented prevented us us from from generatinggenerating reliable reliable iMEP iMEP data. data. Therefore, Therefore, we we had had to to find find alternative alternative ways ways to to analyze analyze the the combustioncombustion characteristics. characteristics.

FigureFigure 10. 10.Overall Overall RLE RLE prototype, prototype, showing showing encoder encoder location location and and revised revised gearbox. gearbox. For For more more details details onon the the prototype prototype geometry, geometry see, see the the Appendix AppendixA .A.

We present a number of BSL cases later in the paper. Only one of them has high resolution cylinder pressure data (using the same encoder as the BSL). The rest of the cases, taken more recently, do not use the encoder because we do not want to upset our TDC alignment on the RLE prior to taking high load data in the future. The high-resolution pressure data for the BSL are for the high compression ratio case that was presented by Appl. Sci. 2021, 11, x FOR PEER REVIEW 14 of 32

We present a number of BSL cases later in the paper. Only one of them has high Appl. Sci. 2021, 11, 779 resolution cylinder pressure data (using the same encoder as the BSL). The rest 14of of the 31 cases, taken more recently, do not use the encoder because we do not want to upset our TDC alignment on the RLE prior to taking high load data in the future. The high-resolution pressure data for the BSL are for the high compression ratio case that was presented Dardalis et al. (2019) [7]. These data can be seen as the black line in Figure 11, where the by Dardalis et al. (2019) [7]. These data can be seen as the black line in Figure 11, where top end of the P-V diagram is shown. the top end of the P-V diagram is shown.

FigureFigure 11.11. TopTop ofof the the PV PV diagram diagram of of the the original original BSL BSL case case (black), (black), RLE RLE with with original original injection injection timing tim- (red),ing (red), and RLEand RLE with with retarded retarded timing timing (blue). (blue).

3.3.3.3. ComparisonComparison ofof HighHigh CompressionCompression RatioRatio BSLBSL withwith thethe RLERLE InIn thisthis subsection,subsection, wewe showshow somesome P-VP-V diagramsdiagrams thatthat helphelp usus analyzeanalyze thethe combustioncombustion characteristicscharacteristics of the the various various cases cases that that we we are are presenting. presenting. If the If thepressure pressure rise risedue to due com- to combustionbustion has hasthe theslope slope shown shown by the by theblack black and and red redcurves curves of Figure of Figure 11, 11it ,means it means that that the thevolume volume was was still stilldecreasing decreasing when when combustion combustion occurred; occurred; therefore, therefore, the combustion the combustion took tookplace place before before TDC. TDC. In the In case the of case the of blue the curve, blue curve, the pressure the pressure rise is risealmost is almost vertical, vertical, mean- meaninging that thatthere there was was no noappreciable appreciable volume volume change change during during combustion, combustion, so so the premixedpremixed combustioncombustion tooktook placeplace atat thethe TDC.TDC. IfIf thethe slopeslope waswas reversed,reversed, thenthen thethe pressurepressure riserise wouldwould havehave beenbeen takingtaking placeplace whilewhile thethe volumevolume waswas increasing,increasing, andand itit wouldwould havehave takentaken placeplace afterafter TDCTDC (this (this is is the the typical typical way way diesel diesel engines engine ares are set set up up to to run). run). Additionally, Additionally, a portion a portion of theof the top top of theof the compression compression and and expansion expansion curves, curv ases, in as the in redthe red graph graph of Figure of Figure 11, overlaps. 11, over- Thislaps. means This means that the that combustion the combustion had ended had before ended the before minimum the volumeminimum was volume reached, was so thereached, pressure so the followed pressure the followed corresponding the corresponding adiabatic curve adiabatic from curve the end from of combustion,the end of com- to thebustion, minimum to the volume, minimum and volume, back over and the back early ov expansion.er the early In expansion. the black curveIn the of black Figure curve 11, someof Figure degree 11, some of combustion degree of continuedcombustion until cont theinued minimum until the volume minimum or volume slightly or after; slightly the expansionafter; the expansion followed afollowed slightly highera slightly adiabatic higher curveadiabatic than curve the late than compression. the late compression. AsAs discusseddiscussed above,above, thethe RLERLE hashas beenbeen upgradedupgraded withwith aa newnew gearboxgearbox andand alsoalso withwith aa properproper TDCTDC location.location. TheThe datadata ofof FigureFigure 1111 (red(red line)line) werewere takentaken beforebefore everever modifying modifying thethe injectioninjection pump timing, timing, and and with with the the same same injector. injector. Additionally, Additionally, we we are are still still using using the the exact same injection pump as we had in the original BSL experiments. It is obvious exact same injection pump as we had in the original BSL experiments. It is obvious from from Figure 11 that the injection timing in the RLE was substantially over-advanced and Figure 11 that the injection timing in the RLE was substantially over-advanced and all all combustion was complete prior to the TDC, whereas for the BSL, even though the combustion was complete prior to the TDC, whereas for the BSL, even though the pre- premixed combustion was clearly taking place prior to the TDC, some combustion and mixed combustion was clearly taking place prior to the TDC, some combustion and in- increase in pressure continued after the TDC. The RLE red line graph also shows a pressure crease in pressure continued after the TDC. The RLE red line graph also shows a pressure oscillation that was not present in the data presented by Dardalis [7]. It seems that we oscillation that was not present in the data presented by Dardalis [7]. It seems that we are are having some abnormal combustion along with the pressure wave in the channel that having some abnormal combustion along with the pressure wave in the channel that con- connects the combustion chamber to the pressure transducer (a channel of approximately nects the combustion chamber to the pressure transducer (a channel of approximately 20 20 mm is connecting our transducer to the combustion chamber, see Figure A2; under certain conditions, a pressure oscillation is seen in the signal). In the blue case in Figure 11, the injection timing was retarded (still on the RLE). As expected, the peak pressure was reduced, making it a more fair comparison; however, some pressure noise towards the very top is still visible (this is even more apparent on the Pressure vs. Crank Angle graph). Appl. Sci. 2021, 11, x FOR PEER REVIEW 15 of 32

Appl. Sci. 2021, 11, 779 mm is connecting our transducer to the combustion chamber, see Figure A2; under certain 15 of 31 conditions, a pressure oscillation is seen in the signal). In the blue case in Figure 11, the injection timing was retarded (still on the RLE). As expected, the peak pressure was reduced, making it a more fair comparison; however, some pressure noise towards the very top is Thestill visible blue graph (this is in even Figure more 11 appa appearsrent on to the be aPressure much bettervs. Crank comparison Angle graph). to the old BSL dataThe (black). blue graph However, in Figure the combustion11 appears tostill be aappears much better to be comparison faster. We to are the not old exactly BSL sure dataas to (black). why, other However, than the combustion fact that we still were appears forced to tobe servicefaster. We the are injector not exactly between sure the two astests to why, due other to contamination, than the fact that and we the were cracking forced to pressure service the may injector have between increased. the Astwo we said testsabove, due the to obsoletecontamination, mechanical and the fuel cracking system pressure cannot bemay controlled have increased. directly; As instead, we said we tried above, the obsolete mechanical fuel system cannot be controlled directly; instead, we tried different operating parameters in a trial and error mode and observed the results in the different operating parameters in a trial and error mode and observed the results in the pressure signal. pressure signal. We had a new injector prepared with a lower cracking pressure and tested it in the We had a new injector prepared with a lower cracking pressure and tested it in the RLE.RLE. The The result result was was a a slower slower combustion combustion ev eventent (the (the combustion combustion duration duration and and location location with withrespect respect to the to TDCthe TDC is presented is presented later later with with the the instantaneous instantaneous heatheat release rate rate calcu- calculations, ◦ lations,see Section see Section 3.7). Again, 3.7). Again, we are we focusing are focusing on theon the conditions conditions of of 750 750 rpm rpm and and 71 71 C°Ccoolant coolantand oil and temperature. oil temperature. The results The results are shown are shown in Figures in Figures 12a and12a 13andb. 13b. Figure Figure 12a 12a shows the showspressure the vs.pressure crank vs. angle crank graph, angle where graph, the wh slowerere the combustionslower combustion also minimized also minimized the pressure thewave pressure that we wave obtain that on we the obtain signal on rightthe signal at the right end at of the the end premixed of the premixed combustion. combus- Figure 12b tion.shows Figure the top12b ofshows the P-Vthe top diagram of the ofP-V the diag sameram case, of the which same alsocase, lookswhich remarkably also looks re- identical markablyto the BSL identical in Figure to the 11 BSL(which in Figure is the 11 same (which case is as the Figure same 9case for as BSL). Figure Therefore, 9 for BSL). it can be Therefore,concluded it thatcan be the concluded case in Figure that the 12 caseis a muchin Figure better 12 is match a much to better the original match to BSL the cases in originalterms of BSL percentage cases in terms of fuel of percentage burnt at the of TDC.fuel burnt The at comparison the TDC. The of comparison PCP minus of PMP PCP of these minustwo cases PMP is of therefore these two more cases relevant.is therefore more relevant.

Appl. Sci. 2021, 11, x FOR PEER REVIEW 16 of 32

(a)

(b)

FigureFigure 12. 12. (a() aOne) One case case of RLE of RLE combustion combustion characterist characteristicsics that matches that matches well the wellBSL and the BSLlacks andthe lacks the pressure wave noise. (b) Top of P-V diagram for the RLE case, also identical to BSL in Figure 11. pressure wave noise. (b) Top of P-V diagram for the RLE case, also identical to BSL in Figure 11. Unlike Figure 11, the proper clearance volume is now used for the BSL graph. Unlike Figure 11, the proper clearance volume is now used for the BSL graph. Pressure, kPa

(a) Pressure, kPa

(b) Figure 13. (a) Pressure vs. crank angle history on the RLE when the timing is retarded. Note that combustion is still fairly fast compared to BSL Case 3 (see Appendix A), but we can still consider this a close comparable case for PCP-PMP comparison. (b) Pressure vs. volume for the RLE in the Appl. Sci. 2021, 11, x FOR PEER REVIEW 16 of 33

(b) Figure 12. (a) One case of RLE combustion characteristics that matches well the BSL and lacks the Appl. Sci. 2021, 11, 779 pressure wave noise. (b) Top of P-V diagram for the RLE case, also identical to BSL in Figure 11.16 of 31 Unlike Figure 11, the proper clearance volume is now used for the BSL graph.

(a)

(b)

FigureFigure 13.13. (a) (Pressurea) Pressure vs. vs.crank crank angle angle history history on the on RLE the RLE when when the timing the timing is retarded. is retarded. Note Notethat that combustioncombustion is still is still fairly fairly fast fast compared compared to BSL to BSL Case Case 3 (see 3 (seeAppendi Appendixx A), Abut), butwe can we canstill stillconside considerr this a close comparable case for PCP-PMP comparison. (b) Pressure vs. volume for the RLE in the this a close comparable case for PCP-PMP comparison. (b) Pressure vs. volume for the RLE in the retarded condition. Note that now the premixed combustion is right at the TDC with some retarded condition. Note that now the premixed combustion is right at the TDC with some signiﬁcant heat release occurring past the TDC, which lowers PCP. This is to be contrasted against Figure 12b, where the premixed combustion is well before the TDC, and only a slight amount of heat is released before the TDC, or Figure 11 (red), where all the heat is released well before TDC.

The revised estimates for fuel reduction to sustain idle and therefore, friction reduction, are presented in Section 3.6 below. The purpose of the above discussion was to justify the comparison of the BSL Case 1 data to that of the RLE Case 5 data for both heat release and PCP/PMP analysis.

3.4. BSL Testing under Reduced Compression Ratio We decided to conduct a few more tests with the BSL and obtain some preliminary CO2 exhaust concentrations from it. The BSL revival was performed without the crankshaft encoder, since we did not want to move this from the RLE and upset our TDC position. However, we could still take cylinder pressure data via the oscilloscope, which are presented in this subsection. Prior to reviving the BSL, we also decided to reduce its compression ratio, so it would be a better match to the RLE; the correction factor due to differences in clearance volume would hopefully become 1 or closer to 1. As it turned out, the compression ratio of the BSL became slightly lower, so the correction factor was still needed, but the correction factor now ampliﬁed the difference in PCP-PMP rather than reducing it. The variation of compression ratio was performed by replacing the original thick copper head gasket with Appl. Sci. 2021, 11, 779 17 of 31

a stock gasket, but apparently, given that we did not know exactly the extent to which each gasket crushes, the compression ratio did not end up at the exact value of the RLE (we used a copper head gasket in the RLE as well, but the head bolts were torqued down to a lower torque than the BSL, since the combustion gases were contained by our RLE hydrodynamic face seal). The compression ratio comparison was performed via our motoring test. This motoring test is not exactly a motoring test, since we do not have a motoring dynamometer. Instead, we rev the engines up, turn off the injection pump, and record pressure and speed. This can be conducted with the data acquisition system, or simply by the oscilloscope. This process is shown in the AppendixA. The process yielded 28 bar absolute (27 bar gage) at around 1080 rpm (the sensitivity of the peak pressure on rpm is very low for both engines; it used to be slightly more pronounced for the BSL when it operated with a higher compression ratio). The comparison in terms of CO2 concentration is shown in Tables1 and2. Compar- ing the cases with advanced injection timing, the RLE exhibited about 20% lower CO2 concentration for the idle condition in question. With the retarded timing, the difference became substantially wider at a 45% difference. This margin is larger than the pressure data suggest, and we cannot explain this large difference yet.

Table 1. Summary of PCP minus PMP for 750 rpm and 71 ◦C (160F) coolant/oil idle conditions. RLE cases in italics font.

Case Number PCP (bar abs) PMP (bar abs) Difference (bar) 1. BSL high compression (old) but combustion continues after TDC. 59.5 38 21.5 2. BSL low compression, med advance 44.5 28 16.5 3. BSL low compression, retarded timing. 39.0 28 11.0 4. RLE old (combustion completed well before TDC) 49.8 33 16.8 5. RLE new (some combustion taking place past TDC) 45.6 33 12.6 6. RLE new, slightly cooler coolant, further retarded timing. 42.5 * 33 9.5 * In other tests with different injectors, we have seen PCP as low as 40 bar when the temperature approaches 71 C (160F).

Table 2. Summary of overall results plus total heat release and CO2 emissions (for details on heat release, see Section 3.7).

Difference Adjusted to RLE Total Adiabatic Heat % Idle Fuel Economy Improvement Case Number CO (%) Clearance Volume (bar) Release (j/cycle) 2 Based on PCP-PMP 1. 20.5 487 ** N/A 2. 18.6 N/A 2.52% 3. 12.4 N/A 2.54% 4. 16.8 328 2.02% 18% compared to 1, 5. 12.6 319 1.31% 32% compared to 2, 38% compared to 1. 6. 9.5 325 N/A 23% compared to 3 ** described in Section 3.7.

During these tests, it was discovered that some further correlations between PCP-PMP can be made, and the injection timing can be used as a variable. Specifically, we know that as the injection timing is retarded, the PCP will be reduced. We wanted to capture this effect and compare it to how the RLE responds to similar injection timing. This allowed us to group cases with similar combustion characteristics, and make more refined comparisons, thus eliminating the uncertainty of injection timing. Even though we can no longer define the exact position of the premixed combustion with respect to the TDC (given the absence of the crankshaft encoder from the BSL), the changes in the combustion characteristics are fairly large, and solid conclusions can be made, even in the absence of an encoder. In this next series of experiments, the new injector with reduced cracking pressure was used with the BSL as was the case of the RLE data given in Figure 12b. In this case, it has a rather advanced injection timing. This is evident by the fact that the sharp pressure rise Appl. Sci. 2021, 11, 779 18 of 31

of the premixed combustion takes place at a compression pressure of at least 3 bar below the maximum motoring pressure and of course, prior to any local maxima of the pressure. When the temperature actually reached 71 ◦C, the peak pressure was about 44.5 bar abs. The combustion behavior of this case is shown in the AppendixA (Figure A4). This case is called Case 2. Now, let us explore the PCP response when the timing is retarded by 4 degrees. The combustion characteristics are again shown in the AppendixA (Figure A5). The fact that these conditions are of a retarded injection timing can be concluded by the fact that the premixed combustion started at a pressure about identical to PMP, indicating that combustion started around the TDC and most of the heat release took place after TDC. For some reason, the premixed combustion is not as well deﬁned in this case, possibly due to the reduced compression ratio. The combustion seems to be slower than other cases. The PCP for that case was about 39.0 bar absolute. This case, which we call Case 3, is also shown in the AppendixA. Obviously, we cannot use the difference of PCP-PMP to compare this case with, for example, the RLE cases of Figure 11 or Figure 12, since in these cases, most of the heat release was completed by TDC. However, an RLE case with similarly retarded timing is presented in Section 3.5 in order to show that the trend of lower PCP-PMP in the RLE persists.

3.5. RLE Testing under Retarded Injection Timing In order to be consistent with the BSL and cover the entire range of injection timing, the RLE was also retarded by 4 degrees from the case of Figure 12. Figure 13a shows the resulting pressure vs. CA (Crank Angle) history. It is clear that the premixed combustion starts just before TDC and ends just after TDC. Unlike the BSL, however, we have a well-deﬁned premixed combustion instead of a more gradual pressure rise associated with a small degree of premixed combustion. Due to a coolant leak, we stopped the test early, and this case did not quite reach 71 ◦C, so we are presenting 63 ◦C with a higher fMEP, making the comparison more conservative. The trend is clear. The PCP dropped substantially, as in the BSL. Additionally, a slight pressure wave noise returned right around the TDC probably due to the rapid combustion. The top part of the PV-diagram is shown in Figure 13b, where the characteristics of the combustion are again conﬁrmed. Additionally, in both Figure 13a,b, the effects of evaporative cooling by the injected fuel are visible, creating a local reduction in the pressure just before combustion.

3.6. The PCP-PMP Comparison Method, and First Estimate of RLE Friction Reduction The above cases set the stage for meaningful comparisons of PCP minus PMP of comparable RLE and BSL cases in order to establish the total heat release needed to sustain the engine. It can be assumed that the indicated thermal efﬁciency of both engines is about identical, and since the only load on the engine at idle is its own internal friction, the fMEP is directly proportional to the heat needed to sustain the engines. This total heat release per cycle is proportional to the PCP-PMP, as long as the comparison is made for cases where about the same percentage of the total heat release is complete at the TDC, and/or as long as the peak cycle pressure is around the TDC (in other words, as long as the combustion phasing with respect to the TDC is identical). The only case where peak pressure is substantially after the TDC is Case 3 (see Figure A5 in the AppendixA). We collected enough information to justify meaningful comparisons. For the cases shown in Tables1 and2, we can compare Cases 1 and 4 (favorable for the BSL as discussed), 1 and 5, 2 and 5, and 3 and 6. The CO2 measurements are not complete, but appear consistent with the PCP-PMP comparison method. The most important corroboration is actually via the integrated adiabatic heat release calculations, which will be discussed in Section 3.7; they show where in the cycle the heat release is taking place. Overall, the comparison in Tables1 and2 is consistent with the trends. The lowest margin of RLE improvement in fuel economy is when comparing Cases 1 and 4, which is 18%. RLE Case 4 is considerably over-advanced as shown by its P-V diagram (Figure 11, Appl. Sci. 2021, 11, 779 19 of 31

red curve). This over-advancing was also apparently causing excessive heat loss, and the actual indicated thermal efficiency was low, as indicated by the higher (compared to RLE Case 5) exhaust CO2 concentration. However, when the timing was retarded one step (Case 5), and the combustion of the RLE was a lot closer to BSL Case 1, we obtained a 38% reduction in PCP-PMP compared to Case 1. However, we are comparing the RLE with a higher compression ratio BSL, where it can be argued that the pressure loading on the BSL piston rings creates an excessive fMEP loss for the BSL (the RLE piston rings and skirt are operating hydrodynamically, so their fMEP contribution is not sensitive to pressure loading). However, when we compare RLE Case 5 with BSL Case 2, where the compression ratio is actually lower than the RLE, we are still obtaining a 32% reduction in fuel consumption. When the timing is considerably retarded on both engines, the margin is reduced to 23%. This is reasonable, based on the fact that the retarded timing reduces the pressure loading on the BSL piston rings, but also, it needs to be reiterated that this comparison favors the BSL where the peak pressure is considerably past the TDC (Figure 13a for the RLE and Figure A5 for the BSL). Some readers may wonder about the validity of the PCP-PMP method, by suggesting that the leakage from the face seal reduces the peak pressure on the RLE. This was expanded upon by Dardalis et al. 2019 [7], but it will be briefly mentioned here. The combustion with this low-pressure injection system is very sooty and leaves heavy traces where there is blowby. When the engine is disassembled and inspected, it is typically observed that the area of the seal is clean from soot, and in fact, the area where the soot never enters in the seal gland is very well defined (See Figure A2 in the AppendixA). This is in contrast to the piston ring lands, for example, where soot is clearly defining the leakage path, especially in the RLE where the rings are pinned, and the end-gaps are always in the same position. Second, the RLE face seal area is directly connected to the gear area (Figure4 or Figure A1), which is in turn isolated from the main crankcase. We had leaks occur occasionally due to malfunctions, caused either by an error during assembly, or on one occasion, by debris from a proximity sensor that collided with the piston and packed in the gland, interfering with the face seal function. When a leak happened, the sound was clearly heard from the open port of cylinder #1. This is the only time when signs of soot are observed in the seal area during disassembly. Finally, a gas leakage is a direct energy loss, which will require a larger quantity of fuel to be burned to compensate. As will be seen in Section 3.7, most of the heat release takes place during a very brief portion of the cycle when there would be very little time for gas leakage. A significant energy loss due to leakage would show itself as a higher amount of total heat released per cycle needed to sustain the engine. However, the overall heat release follows the same trends as the PCP-PMP method, as shown in Section 3.7. Based on the results of this section, it is reasonable to conclude that the RLE fMEP is approximately 20–35% less than the BSL, as it takes about 20–30% less fuel to run, given approximately identical thermal efficiencies. Additionally, given that each of the single cylinders has to rotate a 5-bearing crankshaft, and the oil and water pumps are sized for complete 6-cylinder engines, it can be concluded that the fMEP reduction for a complete engine would likely be much higher than 30%. This is expanded on in Section 3.8.

3.7. Instantaneous and Integrated Total Heat Release Comparison We have high-resolution cylinder pressure data with a crankshaft encoder for many cases for the RLE and one case for the BSL, Case 1 having a high compression ratio. These were analyzed to obtain the instantaneous adiabatic heat release rate. This information can be used to conﬁrm the statements of previous sections with respect to the heat release distribution in the cycle, and speciﬁcally with respect to the TDC. For example, was Case 4 of Tables1 and2 really more advanced than Case 1, and is Case 5 indeed a better match to Case 1 to the position of the combustion event with respect to the TDC? Appl. Sci. 2021, 11, 779 20 of 31

The equation used to calculate the instantaneous heat release rate was borrowed from Brunt and Platts [8] and is shown here as Equation (1).

γ 1 dQ = ·p·dV + ·V·dp (1) hr γ − 1 γ − 1

where Qhr is the heat input by combustion, γ is the ratio of specific heats (function of temperature), p is the pressure, and V is the volume. Note that we ignored heat loss; we are only considering the adiabatic heat release in order to be consistent with the prior PCP-PMP method. For all the graphs shown in this section, the dQ plotted is for 1/20th of a crank angle, which is our encoder resolution. Figure 14 shows the heat release rate for BSL Case 1. This confirms the prior statements that most of the heat release was prior to the TDC, but some combustion continued after the TDC. Additionally, it confirms that the bulk of the combustion takes place in the premixed section. Figure 15 describes Case 4 for the RLE, which had a more advanced combustion. This calculation indicates that while the combustion started at about the same time, virtually all the heat was released in the premixed stage, and all prior to the TDC. Note also that the maximum heat release rate around crank angle −8.5 is very high in Case 4, which is responsible for creating the pressure wave noise right at the end of the premixed phase. This pressure wave also causes a local calculated negative heat release rate at about crank angle −6.5, but this noise does not affect the average calculation. The heat release history confirms that the comparisons of PCP-PMP between Cases 1 (BSL, high compression ratio) and 4 (RLE most advanced) were not entirely fair for the RLE because the combustion characteristics of Case 4 will generate a higher PCP for the same total heat released per cycle. The comparison of the total heat released per cycle of 487 J for the BSL to 328 J for the RLE indeed shows a larger margin for the RLE, a reduction in total heat input by 33%, while the simplified PCP-PMP method only yields 18%. Figure 16 shows the heat release rates for both RLE Case 5 and Case 6 (retarded yet another 4 degrees). The first lesson from Figure 16 is the effect that the combustion retardation has on PCP. As expected, as the injection timing is retarded, the PCP drops. This will obviously happen on both BSL and RLE. Therefore, in the PCP-PMP method, it is Appl. Sci. 2021, 11, x FOR PEER REVIEW 21 of 32 important to compare cases with similar injection timing as expressed by the location of the heat release with respect to the TDC.

FigureFigure 14.14. Instantaneous heat heat release release rate rate for for BSL BSL Case Case 1 1along along with with the the corresponding corresponding cylinder cylinder pressure.pressure. TheThe premixedpremixed combustioncombustion lookslooks sharpersharper thanthan whatwhat isis normallynormally reportedreported inin thethe literature,literature, partlypartly duedue toto thethe low-pressurelow-pressure fuelfuel injectioninjection andand partly partly due due to to the the over-advanced over-advanced timing. timing.

Figure 15. Instantaneous heat release rate for RLE Case 4.

Appl. Sci. 2021, 11, 779 Figure 14. Instantaneous heat release rate for BSL Case 1 along with the corresponding21 of cylinder 31 pressure. The premixed combustion looks sharper than what is normally reported in the literature, partly due to the low-pressure fuel injection and partly due to the over-advanced timing.

Appl. Sci. 2021, 11, x FOR PEER REVIEW 22 of 32

The simplified PCP-PMP method yields a 38% reduction in fueling, while the integrated FigureFigureheat release 15. 15.Instantaneous Instantaneous comparison heat heat (319 release release vs. rate 487 rate for J) RLE yieldsfor RLE Case 34.5%. Case 4. 4.

Figure 16 shows the heat release rates for both RLE Case 5 and Case 6 (retarded yet another 4 degrees). The first lesson from Figure 16 is the effect that the combustion retardation has on PCP. As expected, as the injection timing is retarded, the PCP drops. This will obviously happen on both BSL and RLE. Therefore, in the PCP-PMP method, it is important to compare cases with similar injection timing as expressed by the location of the heat release with respect to the TDC. Now, we consider if, indeed, RLE Case 5 (retarded by about 4 degrees) is a more fair comparison with BSL Case 1. Further examining Figure 16, we note that in Case 5, we can see that the maximum rate of combustion is much reduced compared to Case 4 (Figure 15), causing a reduction in the pressure wave noise, which also reduced the negative heat release just after the premixed combustion, and that some combustion continues after the TDC, much like Case 1 of the BSL shown in Figures 14 or 11. The phasing of the maximum heat release rate, the beginning and end of the premixed combustion (starts at −10, ends at −5), and the small tail end of heat release that stretches to just after the TDC, all match FigurebetweenFigure 16. 16.Instantaneous theInstantaneous two cases. heat heat Case release release 4, rate in rate forcontrast RLEfor RLE Case (Figure Ca 5 (blue/green)se 5 (blue/green)15), has and a Case muchand 6Case (red/purple). higher 6 (red/purple). maximum One rate interestingofOne heat interesting release, observation observation and inthe Figure premixed in 9Figure is that 9 theeventis that approximately wasthe approximately just about 4 degrees over 4 degrees of physicalat −7.5 of physicalcrank rotation angle ofrotation the degrees. of mechanicalThisthe mechanical confirms injection injectionthat pump Cases pump generated 1 and generated 5 about are better 6about degrees matched6 degrees of delay offor in delay the the premixed PCP-PMPin the premixed combustion method combustion event. comparison. event. Now, we consider if, indeed, RLE Case 5 (retarded by about 4 degrees) is a more fair comparisonIf we compare with BSL CaseRLE Cases 1. Further 4 (over-advanced) examining Figure and 16 ,5, we given note that inwe Case have 5, identical we can in- seetegrated that the heat maximum release rate(heat of per combustion cycle), one is much would reduced expect compared that the CO to2 Case emissions 4 (Figure would 15), also causingbe about a reductionidentical. However, in the pressure Tables wave 1 and noise, 2 show which a large also difference reduced the in CO negative2 concentration. heat releaseThis discrepancy just after the is premixed caused by combustion, the fact that and our that heat some release combustion calculation continues is adiabatic after the and TDC,does muchnot consider like Case the 1 loss of the in BSLindicated shown thermal in Figure efficiency 14 or Figure caused 11 .by The the phasing excess heat of the loss of maximumCase 4 (in heatCase release 4, we know rate, the that beginning the heat andloss endis greater of the due premixed to thecombustion longer exposure (starts of the at −10, ends at −5), and the small tail end of heat release that stretches to just after the combustion chamber walls to high temperature and pressure). In other words, our analy- TDC, all match between the two cases. Case 4, in contrast (Figure 15), has a much higher sis only considers the heat release that was directly used to raise the cylinder pressure and maximum rate of heat release, and the premixed event was just about over at −7.5 crank perform mechanical work, and not the heat lost to the combustion chamber walls. The angle degrees. This conﬁrms that Cases 1 and 5 are better matched for the PCP-PMP methodindicated comparison. thermal efficiency The simpliﬁed with respect PCP-PMP to methodexpansion yields ratio, a 38% however, reduction is identical, in fueling, which whileis consistent the integrated with the heat similar release values comparison to the to (319tal adiabatic vs. 487 J) yieldsheat releas 34.5%.e. We can differentiate, however,If we comparebetween RLEthe friction Cases 4reduction (over-advanced) and fuel and consumption 5, given that reduction. we have The identical friction re- integratedduction is heat likely release directly (heat proportional per cycle), one to would the adiabatic expect that total the heat CO2 emissionsrelease. However, would also the re- beduction about identical.in fuel consumption However, Tables at idle1 and will2 show be proportional a large difference to the in total CO 2heatconcentration. release (which Thisincludes discrepancy the heat is lost caused to the by combustion the fact that cham our heatber walls). release Brunt calculation and Pratts is adiabatic [9] suggest and that the actual heat release is about 15% higher than the adiabatic heat release. Therefore, the percentage reduction in fuel consumption may prove to be 15% higher (i.e., a 30% value would be 34.5%). We are planning on performing fuel flow measurements in the future. Given that the PCP-PMP method is corroborated by the more sophisticated heat release rate calculation, we can also compare Case 5 with Case 2 (28% benefit for the RLE), as well as Case 3 with Case 6 (20% benefit for the RLE even though Case 6 is not entirely warmed up). It is only reasonable for the RLE margin to be reduced when the timing is retarded, as the pressure loading on the BSL piston assembly is relieved (but also, the RLE was not fully warmed up). However, even with a reduced compression ratio for the BSL compared to the RLE, the substantial friction reduction persists. The trend of greater fMEP reduction at increasing pressure loading (i.e., higher loads) will likely persist (based also on the SVE experience), but unfortunately, the testing precision will have to be improved, since a smaller fraction of the fuel energy will be expended towards overcoming internal friction. A sensitivity analysis of this heat release calculation was conducted to compare the estimated magnitude of the friction reduction of the RLE at idle with the potential error. Appl. Sci. 2021, 11, 779 22 of 31

does not consider the loss in indicated thermal efficiency caused by the excess heat loss of Case 4 (in Case 4, we know that the heat loss is greater due to the longer exposure of the combustion chamber walls to high temperature and pressure). In other words, our analysis only considers the heat release that was directly used to raise the cylinder pressure and perform mechanical work, and not the heat lost to the combustion chamber walls. The indicated thermal efficiency with respect to expansion ratio, however, is identical, which is consistent with the similar values to the total adiabatic heat release. We can differentiate, however, between the friction reduction and fuel consumption reduction. The friction reduction is likely directly proportional to the adiabatic total heat release. However, the reduction in fuel consumption at idle will be proportional to the total heat release (which includes the heat lost to the combustion chamber walls). Brunt and Pratts [9] suggest that the actual heat release is about 15% higher than the adiabatic heat release. Therefore, the percentage reduction in fuel consumption may prove to be 15% higher (i.e., a 30% value would be 34.5%). We are planning on performing fuel flow measurements in the future. Given that the PCP-PMP method is corroborated by the more sophisticated heat release rate calculation, we can also compare Case 5 with Case 2 (28% benefit for the RLE), as well as Case 3 with Case 6 (20% benefit for the RLE even though Case 6 is not entirely warmed up). It is only reasonable for the RLE margin to be reduced when the timing is retarded, as the pressure loading on the BSL piston assembly is relieved (but also, the RLE was not fully warmed up). However, even with a reduced compression ratio for the BSL compared to the RLE, the substantial friction reduction persists. The trend of greater fMEP reduction at increasing pressure loading (i.e., higher loads) will likely persist (based also on the SVE experience), but unfortunately, the testing precision will have to be improved, since a smaller fraction of the fuel energy will be expended towards overcoming internal friction. A sensitivity analysis of this heat release calculation was conducted to compare the estimated magnitude of the friction reduction of the RLE at idle with the potential error. As discussed above, it was calculated that BSL Case 1 had approximately 4 cc lower clearance volume (the relatively large difference in peak pressure was partly due to valve timing and excessive valve lash that caused an early intake valve closing, plus slightly different camshaft specs). Additionally, there is approximately +/−0.25 crank angle error on the TDC of the BSL for Case 1, plus some errors in the real piston location as opposed to the encoder position, due to the torsional vibrations mentioned earlier (these are more pronounced in the RLE, where the liner driving mechanism adds a torsional spring and flywheel upstream of the encoder, see Figure 10). Therefore, we did a test calculation of the BSL total heat release, assuming extreme errors of 1 full degree TDC shift (towards the direction where the heat release would decrease), plus a reduction in clearance volume of 10 cc instead of 4 cc, which would also reduce the total heat released per cycle (if the clearance volume was that much lower in the BSL, we would not have had reduced the motoring pressure via the minor changes performed prior to resurrecting the BSL unit). Even at this extreme range of parameters, which is far outside the confidence level, the BSL heat release is still 396 J per cycle, or 24% higher than RLE Case 5. The main reason why the heat release calculation is insensitive to the TDC location is that even in Case 1, most of the heat release occurred substantially before TDC. If the premixed combustion took place near TDC, then it is very likely that the calculation would show more sensitivity to TDC location and clearance volume. In other words, given the nature of combustion for both engines for most cases, the integrated heat release calculation is equivalent to the PCP-PMP method. The margin of error on the simpler method is obviously very low as the experimental error is very low compared to the difference. Again, it is unfortunate that we do not have an electronic fuel injection system so that we can directly control the fuel injection rate and timing. However, this heat release rate analysis explores the direct results of fuel injected, and therefore, more than compensates for the above defect. A lot of trial and error is required to obtain comparable cases, but we have proven that we have accomplished that. Our methods may lack precision, but the margins calculated are substantially higher than the margin of error. Appl. Sci. 2021, 11, x FOR PEER REVIEW 23 of 32

As discussed above, it was calculated that BSL Case 1 had approximately 4 cc lower clearance volume (the relatively large difference in peak pressure was partly due to valve timing and excessive valve lash that caused an early intake valve closing, plus slightly different camshaft specs). Additionally, there is approximately +/−0.25 crank angle error on the TDC of the BSL for Case 1, plus some errors in the real piston location as opposed to the encoder position, due to the torsional vibrations mentioned earlier (these are more pronounced in the RLE, where the liner driving mechanism adds a torsional spring and flywheel upstream of the encoder, see Figure 10). Therefore, we did a test calculation of the BSL total heat release, assuming extreme errors of 1 full degree TDC shift (towards the direction where the heat release would decrease), plus a reduction in clearance volume of 10 cc instead of 4 cc, which would also reduce the total heat released per cycle (if the clearance volume was that much lower in the BSL, we would not have had reduced the motoring pressure via the minor changes performed prior to resurrecting the BSL unit). Even at this extreme range of parameters, which is far outside the confidence level, the BSL heat release is still 396 J per cycle, or 24% higher than RLE Case 5. The main reason why the heat release calculation is insensitive to the TDC location is that even in Case 1, most of the heat release occurred substantially before TDC. If the premixed combustion took place near TDC, then it is very likely that the calculation would show more sensitivity to TDC location and clearance volume. In other words, given the nature of combustion for both engines for most cases, the integrated heat release calculation is equivalent to the PCP-PMP method. The margin of error on the simpler method is obviously very low as the experimental error is very low compared to the difference. Again, it is unfortunate that we do not have an electronic fuel injection system so that we can directly control the fuel injection rate and timing. However, this heat release rate analysis explores the direct results of fuel injected, and therefore, more than compensates Appl. Sci. 2021, 11, 779 23 of 31 for the above defect. A lot of trial and error is required to obtain comparable cases, but we have proven that we have accomplished that. Our methods may lack precision, but the margins calculated are substantially higher than the margin of error.

ComparingComparing BSL BSL Case Case 3 3 and and RLE RLE Case Case 5 5 based based on on the the CO CO22 emissionsemissions yields yields a a larger larger marginmargin for for the the RLE. RLE. We We cannot cannot explain explain the the variation, variation, (other (other than than the the possibility possibility that that the the increasedincreased adiabatic adiabatic heat heat release release also also generate generatess increased increased heat heat loss loss and and amplifies amplifies some some of of thethe fuel fuel consumption consumption of of the the BS BSL,L, but but this this is is not not sufficient sufficient to to explain explain the the difference). difference). We We tooktook the the CO CO2 2numbersnumbers twice twice for for both both cases. cases. Given Given so someme variations variations in in the the valve valve timing timing and and volumetricvolumetric efficiency efficiency of of the the two two engines, engines, we we tend tend to to believe believe the the pressure pressure data data more. more.

3.8.3.8. Extrapolation Extrapolation to to a aComplete Complete Engine Engine InIn this this section, section, we we will will describe describe a friction a friction energy energy accounting accounting scenario, scenario, in order in order to ex- to trapolateextrapolate thethe potential potential reduction reduction in inthe the idle idle fuel fuel economy economy of of a acomplete complete 4BT, 4BT, if if we we had had indeedindeed converted converted all all cylinders, cylinders, and and we we were were comparing comparing the the 4-cylinder 4-cylinder RLE RLE to to a a complete complete 4-cylinder4-cylinder baseline baseline engine. engine. Fi Figuregure 17 17 shows shows the the conceptual conceptual CAD CAD of of a amulti-cylinder multi-cylinder RLE. RLE. InIn this this CAD CAD representation, representation, all all the the design design fe featuresatures of of the the rotating rotating liner liner components components as as fabricatedfabricated for for this this prototype prototype are are reproduced reproduced so that so that the theperformance performance characteristics characteristics of this of prototypethis prototype would would be perfectly be perfectly duplicated duplicated without without any changes, any changes, and would and would fit in the ﬁt exist- in the ingexisting block. block.

Figure 17. CAD representation of the multi-cylinder RLE. The critical components are identical to the current prototype, and the model of the block is the production engine, so a direct modiﬁcation is feasible.

Table3 shows a likely percentage distribution of component friction for the original Cummins 4BT, and how the RLE reduces it, in order to approximate the measured friction reduction via analysis of the combustion characteristics of the two single cylinder engines. The numbers assigned ﬁt general friction distributions that are reported in the literature. Note also, according to a motoring tear down friction test of a diesel engine published by Stanton (2013), at 750 rpm, the piston assembly (plus valves, but the valvetrain is a small contributor) was the dominant source of friction by about 75% of the total, so the Table3 numbers may be a bit conservative concerning the piston assembly’s contribution (note, Stanton’s graph does not properly assign the removal of the cylinder head to the piston assembly; this is a typical error in published tear down tests, which are often performed in the wrong order, Dardalis 2019). The units in Table3 are arbitrary, adjusted to give 100% for the 4-cylinder BSL. In the column that corresponds to the 4-cylinder BSL, the crank 10 + 10 corresponds to 10% for the connecting rod bearings and 10% for the main bearings. The accessories include the injection pump, which will have a substantially reduced power requirement when operating on one cylinder. All injection pump pistons accept fuel during engine operation, but only one piston compresses the fuel against the injector—the other three pistons just return the fuel to the tank without compressing it. The piston assembly contribution for the RLE single cylinder was reduced until the RLE single-cylinder total friction was about 25% reduced from the single cylinder BSL and so as to be consistent with the measurements presented above. The result is a total friction reduction of over 40%. The numbers in Table3 illustrate that the measured 25% friction reduction in the single-cylinder RLE will receive a considerable ampliﬁcation when a complete four-engine is converted. The dramatic piston assembly friction reduction in Table3 from 14.25 to 5.0 is consistent with the tear Appl. Sci. 2021, 11, 779 24 of 31

down presented by Dardalis et al. (2005), where a similarly dramatic friction reduction was recorded for the light-duty RLE under motoring conditions.

Table 3. Hypothetical accounting of friction energy percentages at 750 rpm idle relative to the 4- cylinder BSL. The numbers for the 4-cyl. BSL are adjusted to reasonable numbers from the literature. The numbers for the 1-cyl. BSL are adjusted based on the components removed. The 1-cyl. RLE piston is reduced until the measured margin in friction is reached. Then, the RLE 4-cyl. contains the friction of the added components over the 1-cyl. RLE. This scenario shows a 45% reduction in 4-cylinder idle fuel consumption.

4-Cylinder BSL 1-Cylinder BSL 1-Cylinder RLE 4-Cylinder RLE Piston + valves * 65 14.25 5.0 20 Crank 10 + 10 ** 2.5 + 10 2.5 + 10 10 + 10 Accessories 15 10 10 15 Total 100 36.75 27.5 55 * The two-valve/cylinder design has very low valve lift and low stiffness valve springs, so the contribution is low. ** The crank friction is assumed to be half main bearings, half rod bearings. The fMEP shown by Stanton (2013) is about 105 kPa, but this is a motored condition which is known to yield reduced fMEP. The fMEP of a regular diesel at firing idle at the above coolant temperature is about 137 kPa, based on the friction model by Chen and Flynn (1965) and presented again by Dardalis (2012), updated with coefficients that are a better match for modern engines. The calculation used the mid-range of all coefficients except for the hydrodynamic term, which we used the highest value, justified by the relatively low coolant and oil temperature. The above friction model is semi-empirical and calibrated to match a large number of experimental datasets by Southwest Research Institute. The calibration was performed in the late 1990s when the Cummins 4BT engines were being tested and built. Based on our information, Cummins is currently using a similar model with coefficients of about similar values. If we assume that the total friction energy of the four-cylinder BSL in Table3 is distributed among all 4 cylinders, and that the 100 units of friction energy corresponds to 137 kPa, one can calculate the predicted burden for the single cylinder BSL and RLE. The operating IMEP of the single cylinder BSL would be 137 kPa/100 × 36.75 × 4 = 201 kPa. The IMEP for the RLE would be 137/100 kPa × 27.5 × 4 = 150.7 kPa. Indeed, our IMEP calculation for BSL Case 1 is 208 kPa, while the RLE cases vary between 155 and 180 kPa. As discussed in Section 3.2, the IMEP numbers are not reliable due to the torsional vibrations that upset the volume calculation (likely more so for the RLE with the additional shaft portion that drives the rotating liner), but the order of magnitude IMEP seems to be in general agreement. The RLE seems to reduce the fMEP by about 50 kPa at idle. Since the boundary friction assembly term that the RLE minimizes increases at higher loads, the fMEP reduction will increase at these higher loads. We expect that future research will confirm that. Some researchers may find it difficult to believe that a piston friction reduction method can reduce total engine friction by 50 kPa or more. Part of the reason for this resistance is the many erroneous test results reported by the literature. While usually, the researchers mention that confidence in the absolute numbers of their reported fMEP is low and only the reported trends and margins are useful, nevertheless, readers tend to accept the numbers as given. Stanton et al. (2013) show (once corrected) that the piston assembly-loaded friction, plus unloaded piston assembly friction (boundary oil control rings and the hydrodynamic terms) and the valvetrain contribute 69 kPa for motoring conditions at the lowest rpm range. This graph is reproduced as Figure A6 in the AppendixA. Kim et al. (2005) [ 6], in a similar motoring tear down of the light-duty RLE and BSL engines, showed a friction reduction of 50+ kPa under motoring conditions. These motoring tests are relatively simple, so the confidence level in the absolute numbers is high. Fedden et al. [7] presented two naturally aspirated spark ignition engines, one a conventional poppet valve, DOHC (dual overhead cam), and 4 valves per cylinder with stationary liners, and one a sleeve valve Appl. Sci. 2021, 11, 779 25 of 31

(rotating cylinders close to TDC), both operating at a compression ratio of 6.5:1 and at wide open throttle (shown in Figure A8 in the AppendixA). The sleeve valve unit produced about 100 kPa more BMEP, even at a low engine speed where any possible advantage in volumetric efficiency of the sleeve valve cannot be significant. Note that the pressure loading of the piston rings of these two engines is not too far from our idling diesels, due to their very low compression ratio, even though they were operating at full throttle. Therefore, friction reduction of this order of magnitude, if indeed the boundary friction terms are minimized, is possible and consistent with the prior literature. Additionally, as discussed above, if the boundary friction via liner rotation was not indeed nearly entirely removed, the total friction of the RLE would have been substantially higher than the BSL, simply because the high friction coefficient of metallic contact would have burdened not only the piston axial motion (from about 30 degrees before TDC to about 45 after TDC) but also the rotating liner during that period (which travelled about 6.6 cm during that time period, unlike the piston which only travelled about 2.5 cm). Based on the friction model presented by Dardalis et al. (2012) [5], the friction reduction at idle was expected to be about 25%, but the testing showed that it is 40% or more. Therefore, it may be logical to assume that the 3.5% fuel economy benefit at full load and the 6.8% benefit over the Heavy-Duty FTP derived by the assumptions by Dardalis (2012) may both prove conservative. While the fuel economy percentage improvement will reduce by increasing load because the mechanical efficiency of the baseline naturally increases, the fMEP reduction at increasing load will likely increase, and likely more than the prior modeling suggested. There is a proportionality between the boundary portion of the fMEP that the RLE is reducing/eliminating to the peak cylinder pressure, as the Chenn and Flynn [2] friction model suggests (see Figure A7 in the AppendixA). Furthermore, the 6.8% benefit over the HD FTP (Heavy Duty Federal Test Procedure) was calculated by Dardalis based on data from pre-EGR engines with lower cylinder pressures. Therefore, an RLE applied to a modern heavy-duty truck engine will likely produce much larger benefits than anticipated back in 2012.

3.9. Future Work The future work is going to move in two directions. First, we will connect the single- cylinder test engines to a dynamometer in order to compare friction under ﬁring conditions. Obviously, this will be more challenging, since the engine friction will be a smaller fraction of the total heat release in the engine, and the PCP-PMP method will be harder to apply. Simultaneously, we are working on fabricating a complete four- or six-cylinder RLE that is expected to make IMEP measurement feasible.

4. Conclusions The combustion characteristics of the two single-cylinder engines, the rotating liner unit, and the baseline unit were compared. Based on these comparisons, the fuel energy requirement for adiabatic heat input required to operate the RLE unit, including the energy to drive the liner, is between 25 and 35% lower than the baseline engine, which translates into a mechanical friction reduction of the same amount. The analysis presented has a margin of error far smaller than the difference. If one considers the effects of reduced heat loss to the cylinder walls from the reduced pressure rise needed to keep the engine running, the idle energy beneﬁts may be even higher. If the operating temperature was higher, it is likely that the margin would be higher (a higher oil temperature would reduce all other hydrodynamic terms for both engines, and particularly, the parasitic losses of the rotating liner that are all hydrodynamic). This was accomplished without any special optimization of the piston skirt or ring-pack as described by Dardalis 2012. The reduction in friction is of the order of 50 kPa at idle and will likely grow larger with increasing load, based on the friction model offered by Dardalis [5], which was based on the measurements by Chen and Flynn [2] and has been accepted ever since as the standards for predicting diesel engine friction. The friction reduction fraction compared to a complete engine could be of the Appl. Sci. 2021, 11, 779 26 of 31

order of 40% at these low load conditions. The friction reduction percentage will increase at increasing load because the boundary friction of the BSL increases with increasing load (Chen and Flynn [2]; the effect is illustrated in Figure A7 in the AppendixA). The friction reduction fraction will decrease at increasing speed because the hydrodynamic terms for both engines increase, especially the one of the RLE. Dardalis et al. [5] had calculated the fMEP associated with the rotating liner and its seal at about 3.9 kPa at 2000 rpm engine speed, 600 rpm liner speed, and 10 ◦C oil temperature (these tests were conducted at 71 ◦C). In the past 10–15 years or so, many methods of reducing the boundary piston assembly friction have been developed: new piston ring profiles, plateau honing techniques, DLC ring or cylinder coatings, etc. Our baseline engine certainly did not have any of these features. We had new pistons and piston rings in the engines, and the top rings had the traditional symmetrical profile, while the liner of the baseline was well used but still within spec (the cross hatch on the top of the liner was gone as happens fairly early in the life of any diesel engine, but the hone marks were still visible in the mid portion, whereas most engine testing is performed with fresh liners shortly after break-in). All the above methods have limited usefulness in heavy-duty engines, where the engines are expected to be in use long after the effects of the above technologies have faded away (Dardalis [4,10]). Reducing the oil viscosity has some limited effects on engine friction at low loads only, but at the expense of engine life. According to Carden et al. [9], reducing high-temperature, high-shear viscosity by 36% (from 3.5 × 10−3 Pa*s, probably 15w40, to 2.26 × 10−3 Pa*s, probably 5w30) accomplishes a reduction in friction at low load and 800 rpm of about 5 kPa, and up to 15 kPa at 1400 rpm, while the reduction diminishes to almost nothing for higher loads as the metallic contact intensifies (Carden et al. [9] also explain that the overall fMEP measurements presented were excessively low, and only the differences are important, not the absolute values). Note that if the engine tested by Carden was not a fresh engine and instead was about halfway into its operating life, the benefit may have been even smaller. It is possible that the RLE and its seal will work satisfactorily at the reduced viscosity (the analyses presented by Dardalis et al. [5] were for 3.0 × 10−3 Pa*s viscosity and the seal could handle 180 bar peak pressure). It is also possible that the elimination of metallic contact will be sustained at low viscosity, possibly requiring slightly higher liner speeds. Therefore, the low viscosity benefit could be combined with the RLE friction reduction. Another possibly better option for a high efficiency RLE is to use the 15w40 grade oil plus larger clearances for the pistons and bearings, so the hydrodynamic terms can be reduced without sacrificing the load capacity of the bearings as compared to low viscosity and tighter clearances. This lubricant can be lighter on additives and therefore, easier on the exhaust aftertreatment and of lower cost. Additionally, since there will be less wear debris in the oil from the cylinders and rings, the filtration demands may be reduced. The drain intervals may also be extended since there is less reliance on the additives. Obviously, our parallel claim of much longer engine life cannot be confirmed that easily. The wear reduction that we have proven, however, can only come from elimination of the metallic contact. This will automatically lead to wear reduction. We have not yet proven that at higher loads this mechanism will persist. However, we have proven that the mechanism works with about 50 bar of peak cylinder pressure and 1.4 m/s surface liner speed (which corresponds to the 750-rpm crankshaft speed or 250-rpm liner speed). Under high engine load, the engine speed will also be higher, and the liner surface speed will also be two to three times higher, similar to the Sleeve Valve Engines. There is evidence from Ricardo [11] that the bore wear rate of the sleeve valve engines was one-tenth that of the fixed liner engines of their time, for high BMEP aircraft engines and for industrial diesel engines that operated over tens of thousands of hours and never needed sleeve overhaul (there were no anti-wear or anti-corrosion lubricant additives at the time). It is therefore logical to expect that the dramatic wear reduction will transfer to RLEs, even with light lubricant additive packages, which again could help prolong exhaust aftertreatment, or even reduce the size of the aftertreatment devices. The RLE face seal should not have any appreciable wear either, since it is designed to operate with a relatively high film Appl. Sci. 2021, 11, 779 27 of 31

thickness (Dardalis et al. [12]). The economics of diesel operation may be substantially altered, not only via reducing fuel consumption, but by considerably extending the lives of engines, eliminating the need for on-frame overhauls, reducing the cost of the lubricants, and possibly even extending the life of the aftertreatment devices due to the reduction in anti-wear additives.

Author Contributions: D.D.: Original Draft Preparation, Data Analysis, Experimental Work. M.J.H.: Review and Editing, Supervision, Data Analysis. R.D.M.: Supervision. A.B.: Review and Edition, Data Analysis. All authors have read and agreed to the published version of the manuscript. Funding: This research received no external funding. Conﬂicts of Interest: The authors declare no conﬂict of interest. Appl. Sci. 2021, 11, x FOR PEER REVIEW 28 of 32

Appl. Sci. 2021, 11, x FOR PEER REVIEW 28 of 32 Appendix A Appendix A Appendix A

Figure A1. Detail of the rotating liner during a tear down. The active cylinder is number 2. The Figuredriving gear A1. is identicalDetail in size of to the the rotatinggear on the rotating liner liner during, proving a that tear there down. is sufficient The active cylinder is number 2. The space for all cylinders to be converted. Lubricant flows under the rotating liner flange, through the drivingradial passages gear of the is identicalrotating liner face, in size and back to theto the gearsump. on the rotating liner, proving that there is sufficient space forFigure all A1. cylinders Detail of the to rotating be converted. liner during a Lubricant tear down. The flows active cylinder under is the number rotating 2. The liner flange, through the radial driving gear is identical in size to the gear on the rotating liner, proving that there is sufficient passagesspace for all ofcylinders the rotating to be converted. liner Lubricant face, and flows back under tothe therotating sump. liner flange, through the radial passages of the rotating liner face, and back to the sump.

Figure A2. Detail of the cylinder head modiﬁcation. Note that the seal gland area is completely free from carbon deposits. Appl. Sci. 2021, 11, x FOR PEER REVIEW 29 of 32

Figure A2. Detail of the cylinder head modification. Note that the seal gland area is completely Appl. Sci. 2021, 11, x FOR PEER REVIEW 29 of 32 free from carbon deposits. Appl. Sci. 2021, 11, 779 28 of 31

Figure A2. Detail of the cylinder head modification. Note that the seal gland area is completely free from carbon deposits.

Figure A3. Motoring test on the BS, establishing its motoring pressure. The engine is revved up Figure A3. and the injectionMotoring pump testis turned on the off. BS, As the establishing engine continues its motoring to rotate, pressure. we record Thethe peak engine pres- is revved up and theFiguresure. injection Each A3. division pumpMotoring is is 10 turned bar. test The on off. x theaxis As BS,scale the esta enginegivesblishing us continues the speed its motoring estimate. to rotate, The pressu we speed recordre. here The the is aboutengine peak pressure. is revved Each up divisionand1050 rpm.the isinjection The 10 peak bar. pressure Thepump x axis iswill turned scaledrop slightly givesoff. As usto the 750 the engirpm, speed nebut estimate.continues we are using Theto this rotate, speed figure we hereof about record is about27 the 1050peak rpm.pressure.bar gage Each or 28 division bar absolute. is 10 bar. The x axis scale gives us the speed estimate. The speed here is about The peak pressure will drop slightly to 750 rpm, but we are using this ﬁgure of about 27 bar gage or 1050 rpm. The peak pressure will drop slightly to 750 rpm, but we are using this figure of about 27 28 bar absolute. bar gage or 28 bar absolute.

Figure A4. First case of new BSL data with advanced injection timing, equivalent to Case 2. Pre- mixed combustion starts at 24–25 bar absolute, which is about 2–3 bar below the motoring pressure. This is how we know that the injection is advanced. Note that the peak cylinder pressure in this case is about 55 bar. The engine was not fully warmed up yet. When we reached 70 °C, the reduced friction allowed the peak pressure to drop down to 44.5 bar gage or 45.5 bar absolute.

FigureFigure A4. A4.First First case case of of new new BSL BSL data data with with advanced advanced injection injection timing, timing, equivalent equivalent to Case to2. Case Premixed 2. Pre- combustionmixed combustion starts at 24–25starts at bar 24–25 absolute, bar absolute, which is aboutwhich 2–3is about bar below 2–3 bar the below motoring the motoring pressure. Thispres- sure. This is how we know that the injection is advanced. Note that the peak cylinder pressure in is how we know that the injection is advanced. Note that the peak cylinder pressure in this case is this case is about 55 bar. The engine was not fully warmed up yet. When we reached 70 °C, the about 55 bar. The engine was not fully warmed up yet. When we reached 70 ◦C, the reduced friction reduced friction allowed the peak pressure to drop down to 44.5 bar gage or 45.5 bar absolute. allowed the peak pressure to drop down to 44.5 bar gage or 45.5 bar absolute. Appl. Sci. 2021, 11, x FOR PEER REVIEW 30 of 32 Appl. Sci. 2021, 11, 779 29 of 31 Appl. Sci. 2021, 11, x FOR PEER REVIEW 30 of 32

FigureFigureFigure A5. A5.A5. BSL BSLBSL retarded retardedretarded injection injection timing timing at operating operating temperature, temperature,temperature, or or orCase Case Case 3 3 of of3 Tableof Table Table 1.1. The The 1. The fact fact fact that that these conditions are of a retarded injection timing can be concluded by the fact that the pre- thatthese these conditions conditions are are of aof retarded a retarded injection injection timing timing can can be concludedbe concluded by theby factthe fact that that the premixedthe premixedmixed combustion combustion started started at at a a pressurepressure about identiidenticalcal to to PMP, PMP, indicating indicating that that combustion combustion started started combustionaround the TDC started and at most a pressure of the heat about release identical took to place PMP, after indicating TDC. For that some combustion reason, the started premixed around around the TDC and most of the heat release took place after TDC. For some reason, the premixed thecombustion TDC and is most not ofas thewell heat defined release in tookthis case, place possibly after TDC. due For to the some reduced reason, compression the premixed ratio. combustion The combustion is not as well defined in this case, possibly due to the reduced compression ratio. The isreading not as wellhere deﬁnedis 38 bar in gage, this 39 case, bar possibly absolute. due to the reduced compression ratio. The reading here is reading here is 38 bar gage, 39 bar absolute. 38 bar gage, 39 bar absolute.

Figure A6. Tear down motoring figure copied from Stanton et al. [12]. When the cylinder head was removed, the valve train and the loading on the piston skirt and rings were also removed. The FigureFigurevalve A6. train A6. Tear Tearand down pumping down motoring motoring are of coursefigure ﬁgure fairlycopied sma fromll, and StantonStanton already etet al.partiallyal. [[12].12]. When Whencontained the thecylinder incylinder the removal head head was wasremoved,of removed,the manifolds. the the valve valveThe train graph train and erroneously and the loadingthe loading labels on the on the piston th componente piston skirt andskirt as ringsthe and “cylinde wererings also werer head”. removed. also It removed.also The needs valve The to be remembered that the test shown is under motoring conditions. When the engine is idling and valvetrain train and pumpingand pumping are of are course of course fairly fairly small, sma andll, already and already partially partially contained contained in the removalin the removal of the the PCP is about 20 to 30% higher, this energy loss will increase by about the same amount, while ofmanifolds. the manifolds. The graphThe graph erroneously erroneously labels labels the component the component as the as “cylinder the “cylinde head”.r head”. It also It needs also toneeds be to beall theremembered hydrodynamic that andthe testpumping shown losses is under will remomaintoring about conditions. the same. WhenUnder theload, engine the loaded is idling and rememberedpiston assembly that term the testwill shownof course is undergrow even motoring larger. conditions. The part of the When graph the shown engine as is “Piston idling and& the the PCP is about 20 to 30% higher, this energy loss will increase by about the same amount, while PCPRings” is about represents 20 to the 30% hydrodynamic higher, this energy portion loss of the will piston increase assembly, by about plus the the sameboundary amount, friction while of all all the hydrodynamic and pumping losses will remain about the same. Under load, the loaded thethe hydrodynamic oil control rings and which pumping is most lossesly unaffected will remain by cylinder about pressure. the same. The Under RLE load, in this the paper loaded is be- piston piston assembly term will of course grow even larger. The part of the graph shown as “Piston & assemblylieved to eliminate term will a of large course portion grow of eventhe friction larger. between The part the of two the red graph lines, shown plus the as “Pistonoil control & Rings” Rings”boundary represents friction the term. hydrodynamic portion of the piston assembly, plus the boundary friction of therepresents oil control the rings hydrodynamic which is most portionly unaffected of the piston by cylinder assembly, pressure. plus the The boundary RLE in frictionthis paper of theis be- oil lievedcontrol to ringseliminate which a large is mostly portion unaffected of the friction by cylinder between pressure. the two The red RLE lines, in plus this paperthe oil iscontrol believed boundaryto eliminate friction a large term. portion of the friction between the two red lines, plus the oil control boundary friction term. Appl.Appl. Sci. Sci. 20212021, 11, 11, ,x 779 FOR PEER REVIEW 30 of 3131 of 32

Figure A7.A7. Snapshots takentaken fromfrom Chen Chen and and Flynn Flynn (1965). (1965). On On the the left, left, one one of of the the single-cylinder single-cylinder engines isis shown. shown. Note Note that that there there are are three three compression compression rings rings per cylinder, per cylinder, as opposed as opposed to two piston to two pistonrings for rings more for modern more modern engines (includingengines (including ours). This ours). will reduceThis will the reduce pressure the loading pressure for loading the top for thering, top because ring, thebecause sealing the loading sealing is sharedloading by is three shared instead by three of two instead rings, andof two will rings, reduce and the will metallic reduce thecontact metallic intensity contact and thereforeintensity boundaryand therefore friction boundary of the ring friction pack. Thisof the is ofring course pack. at This the expense is of course of at thean increase expense in of the an hydrodynamic increase in th friction.e hydrodynamic On the right, friction. the graph On the shows right, the the fMEP graph of the shows two singlethe fMEP ofcylinder the two research single engines cylinder as research a function engines of PCP. Theseas a function engines of are PCP. robust, Thes ande withengines a short are crankshaft,robust, and withso they a short do not crankshaft, suffer from so the they torsional do not vibrationssuffer from that the our torsional single cylinders vibrations do. that Therefore, our single IMEP cylin- dersmeasurements do. Therefore, are possible. IMEP measurements Note that as the are peak possible. pressure Note increases, that as thethe FMEPpeak pressure increases increases, as well. the FMEPNote also increases that the as slope well. is almostNote also constant that the for allslope speeds, is almost meaning constant that the for engine all speeds, speedonly meaning slightly that the engine speed only slightly reduces the magnitude of this loss. Note as well that the ER-1 design reduces the magnitude of this loss. Note as well that the ER-1 design has higher fMEP and also higher has higher fMEP and also higher sensitivity to peak pressure. The reasons are first that the ER2 sensitivity to peak pressure. The reasons are ﬁrst that the ER2 has externally driven accessories, and has externally driven accessories, and second that the ER2 has a larger and heavier piston skirt second that the ER2 has a larger and heavier piston skirt that distributes the combustion load better that distributes the combustion load better on the cylinder liner, and also probably generates bet- on the cylinder liner, and also probably generates better consistency on the oil ﬁlm deposited on ter consistency on the oil film deposited on the cylinder liner (skirt metallic contact is known to the cylinder liner (skirt metallic contact is known to scrape off oil and generate local oil starvation scrape off oil and generate local oil starvation conditions). The magnitude of the fMEP shown may conditions). The magnitude of the fMEP shown may be different on modern engines, partly due be different on modern engines, partly due to design differences, and partly due to the single cyl- inderto design nature differences, of these andengines, partly bu duet the to trends the single persist. cylinder Note nature that the of these pressure engines, sensitivity but the trendsterm of persist.0.005 for Note the thatER-2 the is pressurevery similar sensitivity to modern term of en 0.005gines, for which the ER-2 typically is very similarranges to from modern 0.004 engines, to 0.007 whichbased typicallyon the design ranges and from wear 0.004 state to 0.007 of the based engine. on the design and wear state of the engine. Appl. Sci. 2021, 11, 779 31 of 31 Appl. Sci. 2021, 11, x FOR PEER REVIEW 32 of 32

Figure A8. Comparison of an SVE to a poppet valve engine with 4 valves per cylinder, reproduced Figurefrom Fedden A8. Comparison (1938). Both ofengines an SVE operate to a poppetnaturally valve aspirated, engine and with with 4 the valves same per compression cylinder, reproduced fromratio. Fedden Fedden (1938).is not concerned Both engines as to what operate portion naturally of the aspirated,BMEP advantage and with of the the SVE same is due compression to ratio. FeddenfMEP reduction, is not concerned and what portion as to what is due portion to volumetric of the efficiency BMEP advantageeffects, but at of the the low SVE speed is due to fMEP reduction,range, there and cannot what be portion any significant is due todifference volumetric in volumetric efficiency efficiency. effects, but at the low speed range, there cannot be any significant difference in volumetric efficiency. References References1. Miura, A.; Shiraishi K. 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