20056083 81 NUMERICAL INVESTIGATION OF PRESSURE REACTIVE PISTON TECHNOLOGY IN A SPARK IGNITION ENGINE

Wooheum Cho Hyundai Motor Company, Korea

Dohoy Jung, and Dennis Assanis W. E. Lay Automotive Laboratory, The University of Michigan, Ann Arbor, Michigan

The novel, two-piece Pressure Reactive Piston technology providing the advantages of Variable Compression Ratio and rapid response to cylinder pressure changes was developed and numerically simulated in order to investigate its dynamic behavior and effects on fuel efficiency and emissions. The analytical model for flame-combustion chamber interaction was newly developed incorporating the VCR feature of the PRP engine and implemented into the engine cycle simulation program. The cycle simulation was subsequently used to explore the potential of the PRP engine over the full operating range and to investigate the effect of the PRP spring set characteristics on engine performance and emissions.

Keywords: Variable Compression Ratio, Pressure Reactive Piston, Brake Specific Fuel Consumption

Introduction sensitive to disturbances or wear to be used in production applications. Evenmore, many VCR technologies require In a typical Spark Ignition (SI) engine, better new engine architecture to implement, which require efficiency and higher performance can be achieved with significant investment to produce. Brevick patented the higher Compression Ratio (CR). However, higher Pressure Reactive Piston (PRP) technology (Patent cylinder pressure associated with higher CR can increase #5,755,192), which is two-piece piston with a spring set the end gas temperature and cause knocking at high load. between the upper and lower pistons. This mechanism In an effort to achieve higher CR while still avoiding with acceptable cost and complexity effectively limits the knocking, various types of Variable Compression Ratio peak cylinder pressures at high loads, while allowing the (VCR) technologies have been developed. engine to operate at high compression ratios under low In general, compression ratio can be varied by loads. changing both the clearance volume and the displaced In this paper, an analytical method for computing volume or only the clearance volume, and the designs flame front propagation has been newly developed for the achieving VCR can be divided as followings: PRP engine in order to model the geometric interaction Eccentric Crankshaft: A lengthened crank-arm not between the spherically propagating flame and the only increases the stroke length and thus the engine changing boundaries in the PRP combustion chamber. In displacement but also decreases the clearance volume [1]. addition, the upper to lower piston motion of the PRP is Linkage System: An additional linkage on the crank implemented into the cycle simulation program and mechanism can be superimposed in order to adjust proper parametric study of the PRP engine is subsequently CR [2~3]. carried out in order to investigate the effect of engine Axial Engine: The angle of a wobble plate is varied speed, load, and spring preload on the PRP engine with respect to its rotational axis or adjusted performance and NO emission. longitudinally to change the piston stroke or clearance volume [4]. Pressure Reactive Piston Configuration Variable Combustion Volume (VCV): The complete cylinder head or a section is adjustable [5~8]. It is The PRP assembly is separated into an upper piston classified further by the control device; Chamber Tilting, or crown and lower piston or skirt by a spring set as Liner Lifting, and Auxiliary Cylinder. shown in Fig. 1. Both piston crown and piston skirt are Variable Compression Height (VCH): The upper made of an aluminum alloy with the spring seat area portion of piston is adjustable [9~11]. It is classified by anodized. Belleville spring geometry was selected as the the piston configuration; Two-Piece Piston with hydraulic spring set for the PRP because of its compactness and device or load-deflected spring and Flexible Piston ability to carry high load with small deflections. A Crown. retainer ring is inserted into the groove of piston crown Variable Radius Length Engine (VRLE): The inner side and functioned to hold the two-piece piston complete piston is controlled by adjusting the radius of together. the connecting rod big end [12~13]. The cycle simulation program, Spark Ignition Simulation (SIS), is used in this study to examine the On the whole, the early VCR designs for existing maximum CR without knocking and provide the engine architecture were technically too complex or too boundary conditions for spring design. In order to

Presentation No. 81 Speaker Name: W. CHO 1/6 20056083 investigate the geometry and configuration of the spring Geometric Model for Flame Propagation set, a spreadsheet program for calculating the relation between spring deflection and force [14] was developed The practical importance of the geometric and stress analysis was performed for reliable and durable interaction between a propagating flame and the spring design. As the result, the CR of the PRP is varied combustion chamber is that the entrained mass rate in the from 9.26 to 13.5, which sets maximum spring deflection quasi-dimensional combustion model is proportional to as 3.3 mm. The spring set is designed to begin to deflect the spherical flame front area and heat transfer between at 22 bar of preload and to be fully defected at 66 bar of the burned gases and the walls is proportional to the peak cylinder pressure. chamber surface area wetted by the burned gases. Most researchers [18~19] modeled the combustion chamber with simple geometry for the calculation of the flame front area and wetted wall area, but this method is not applicable to a combustion chamber with more Upper Piston Lower Piston complex geometry. In particular, for the case of the PRP, it is important to calculate the combustion chamber volume with respect to crankangle, because the PRP has variable compression height depending on engine load. Consequently, Belleville Spring Retainer Ring modeling of the flame front interaction with the combustion chamber of the PRP engine is an important Fig. 1 Pressure Reactive Piston Cross Section pre-requisite for performing simulation studies. The propagating flame in this cycle simulation is assumed to Dynamometer test of PRP Engine develop as a sphere with its center fixed at the spark plug electrodes and truncated by the combustion chamber wall. The experiments of PRP engine were done by using The configuration and geometric notation of the dynamometer at 2000 rpm with various loads. The best combustion chamber used in this research are illustrated improvement in Brake Specific Fuel Consumption in Fig. 2 (a). (BSFC) of PRP engine over baseline was observed as ℓ r 1 s ℓ2

∆ Spark H

7.8 % at light load. This improvement was less than PRP Z CRH Z 21 11 θ C z θ rf θ1 2 expected because the target spring set load/ deflection tdc Z h characteristics were not achieved with the first spring set. S A A ∆ H The further details of experimental results can be found in base Cho [15]. Therefore, the simulation work was performed to find the full potential of the PRP engine and to investigate the effect of preload on fuel economy. The ℓ 6 R ℓ Ricardo Hydra SI engine used in this research is the 5 ℓ4 ℓ3 r θ γ 11R β baseline of this research and its specifications are shown α θ in Table 1. 22R R r R-rs s Engine Type 4-Stroke, Single Cylinder, PFI Bore x Stroke 80.26 x 88.90 mm (a) Hydra Engine (b) Simple Disk Type Connecting Rod 158.01 mm Fig. 2 Geometric Notation Combustion Chamber Compression Ratio 9.26 IVO/IVC 11 deg. BTDC/49 deg. ABDC Flame Front Area & Entrained Volume EVO/EVC 49 deg. BBDC/11 deg. ATDC Rated Power 20 kW @ 5000 rpm As shown in Fig. 2 (b) the flame front area and Table 1 Specifications - 4 Valve Ricardo Hydra SI Engine entrained volume of a simple disk type combustion chamber are calculated from the following equations. Spark Ignition Simulation π 2 2π h 2π h The quasi-dimensional computer simulation of the SI Ae = ∫∫rf cosθdα ⋅ rf dθ =∫∫rf dαdz =2rf ∫αdz (1) engine working cycle is used due to its overall predictive 0 0 0 0 0 ability, particularly regarding the effects of combustion h ⎛ h ⎞ ⎜ 2 2 ⎟ chamber shape on flame propagation. The basis for this V f = ∫ Ac dz⎜= ∫ (αr + βR − rs Rsinβ )dz⎟ (2) work is the quasi-dimensional SI engine cycle simulation, 0 ⎝ 0 ⎠ SIS, which was developed by Poulos and Heywood [16] where Ae is flame front area, Vf is entrained volume, h is and extended by Filipi and Assanis [17]. The combustion piston crown distance from spark plug, rs is the spark sub-model is based on the turbulent-flame entrainment plug distance from the cylinder centerline, and R is model proposed by Tabaczynski [18] and further refined cylinder radius. The entrained volume and flame front by Poulos and Heywood. The earlier model is also area in Eqs. (1) and (2) are calculated with the angle, α, complemented by a zero-dimensional turbulence model and the cross sectional area, Ac, respectively. which calculates average turbulent flow field parameters Based on this result, the first step to calculate the throughout the whole cycle. entrained volume and flame front is that the flame radius

Presentation No. 81 Speaker Name: W. CHO 2/6 20056083 is divided into 11 ranges due to the geometric complexity r + z / tanθ of combustion chamber. Then, for a given crankangle and γ = θ −θ −1 s 1 L d12 d11 θ d11 = cos ( ) flame radius, α and Ac are obtained. Finally, the R integration is done by summing all the α’s and Ac’s over if (R 〈r ),θ = 0 h. Further details on this calculation procedure can be 0 min d12 found in Cho [15]. else _ than _ if (R 〈r ) 0 max Flame Wetted Areas r 2 + R2 − R 2 π θ = cos−1( s 0 ), 1. Combustion Chamber Head d12 2⋅r ⋅ R 2 s θd11 The combustion chamber consists of head, cylinder 2 2 rmax = R − rs wall and piston and the combustion chamber head is θd12 divided into oval, side, and cylinder wall parts. The only R0 R dependent variable for calculation of the combustion r = r 2 + R 2 − 2 ⋅ R ⋅ r ⋅ cosθ chamber head is flame radius. min s s d11 The oval and side area of combustion chamber head 1 R = (r 2 − z 2 ) 2 consists of left and right sides separated from the spark 0 f plug. At each side, the flame radius is divided into 4 ranges for more accurate calculation as shown in Fig. 3. Fig. 4 Notation and Calculation Equation for Wetted Area of the Combustion Chamber Head (Side Part) 0 ≤ rf 〈rh11 rh11 ≤ rf 〈rh12 2. Cylinder Wall

x11 2 π π The wetted area of cylinder wall in Eq. (3) can be S = r S = r 2 − (r 2 x − r 2scx ) h,1L1 f 2 h,1L2 f 2 f 11 f 11 calculated based on peripheral integration. The interval, θ1 is determined at the TDC surface plane as shown in

2 Eq. (4). The wetted distance, h, is the vertical flame ⎛ (r / cosθ )2 + r 2 − R ' ⎞ ⎛ ( / cosθ ) ⎞ θ = cos −1⎜ s 1 f 1 ⎟ x = cos −1 ⎜ l 1 1 ⎟ distance from TDC surface and calculated as Eq. (5). L31r ⎜ ⎟ 11 ⎜ ⎟ 2⋅(rs / cosθ1 )⋅rf r θ1 ⎝ ⎠ ⎝ f ⎠ if {(r 2 − r 2 )0.5 − H}〈S (3) 2 '2 2 A = 2 hdz =2 hRdθ f 0 ⎛ (r / cosθ ) + R − r ⎞ −1 ⎛ (rs / cosθ1 ) ⎞ cw −1⎜ s 1 1 f ⎟ ⎜ ⎟ ∫ ∫ θ = cos θ L3 = cos 0 L31R ⎜ ' ⎟ ⎜ ' ⎟ 2⋅(rs / cosθ1 )⋅ R1 ⎝ R1 ⎠ ⎝ ⎠ (R2 + r 2 − R 2 ) ⎛ ⎞ θ = cos−1{ s 0 } (4) −1 (rs / cosθ1 ) + (l1 / cosθ1 ) 1 θ L4 = cos ⎜ ⎟ 2Rrs ⎜ R ' ⎟ ⎝ 1 ⎠ 2 2 0.5 h = (rf − r0 ) − H (5) rh12 ≤ rf 〈rh13 2 2 2 0.5 θL31R where S = l + a − a cosθ − (l − a sin θ) , dz = Rdθ , θ θL3 L31r r ≤ r 2 2 0.5 2 2 0.5 θL4 h13 f R0 = (rf − H ) , and r0 = (R + rs − 2 ⋅ R ⋅ r ⋅s cosθ ) ,

'2 '2 '2 ' Sh,1L4 = R1 θ3 − (R1 θL4 − R1 scθL4 ) − R1 ⋅rs sinθ3 / cosθ2 H is the distance from the spark plug to the crown 2 2 2 2 π ′ ′ ′ rs ′ position at TDC. The geometric notation is illustrated in Sh,1L3 = rf (θ L31r − ) + R1 θ L31R − (R1 θ11R − R1 scθ11R ) − R1 sinθ L31R 2 cosθ2 Fig. 5 (a).

Fig. 3 Classification and Calculation Equation for Wetted Area of the Combustion Chamber Head (Oval Part)

h h H As Fig. 4 illustrates the geometric configuration, the A 0 A B side area is calculated by the integration of cylinder B S radius, R, multiplied by wetted angle, γ. For a given flame radius, the wetted side area of combustion chamber head dz R R A – A RO is calculated as the following basic steps. First, minimum R0 Section θ2 θ1 and maximum radii are calculated at a cross sectional h θ1 rs plane of a given height, z, from spark plug. The flame h0 radius, Ro, on this plane is calculated based on purely geometric correlation. Comparing these radii and the wetted angle, γ is calculated at both sides. Finally, the B – B r Section integration from the spark plug is done by summing the 0 : θ peripheral circular arcs up to the propagating flame plane. The end point of integration is limited to the piston crown position at TDC. The wetted cylinder area of the combustion chamber head is calculated from the same procedure applied for the cylinder wall area, which will (a) Cylinder wall (b) Piston be explained as follows. Fig. 5 Geometric Notation

Presentation No. 81 Speaker Name: W. CHO 3/6 20056083 3. Piston the figures in order to give clear comparison and visual advantages of the PRP engine. Because the first spring set The surface area of the piston is wetted by the flame is softer than the proposed design and the total deflection only if the flame radius is larger than a distance between is 3.8 mm, the lower CR boundary for the PRP engine the spark plug and piston crown. In this case the wetted with the first spring is not 9.26, but 8.825. Note that a 3.3 piston area is calculated by following equations. mm spring deflection is corresponding to CR=9.26. On A = θ R2 +θ R 2 − r Rsinθ (6) the other hand, a conventional engine with CR=9.26 p 1 2 0 s 1 corresponds to the lower boundary of the PRP engine where with the proposed spring set, while the upper boundary 2 2 2 2 2 2 −1 (R + rs − R0 ) −1 (Ro + rs − R ) corresponds to a conventional engine with CR 13.5. θ1 = cos ,θ2 = cos . 800 2Rrs 2Rors 40 PRP_present Fig. 5 (b) indicates the geometric parameters used in 30 8.825_BASE 650 calculating the equation. 13.5_BASE 20 500 350

Implementation of PRP Motion into SIS (Nm) Torque 10

BSFC (g/(kW-hr)) 200 The piston position of the PRP engine is determined 0 0.2 0.4 0.6 0.8 1 0 300 600 900 1200 by adding the spring deflection to a distance induced by MAP (atm) BMEP (kPa) rotation of crankshaft. A new crankangle corresponding (a) Torque (b) BSFC to this piston position is calculated by using the relation between rotational angle of the crankshaft and normal 2400 344 piston position and the main program searches the flame 1800 340 radius with known values of this new crankangle and 1200 336 enflamed volume. Once the flame radius is obtained, flame front area, entrained volume and wetted areas NO (ppm) 600 332

corresponding to the corrected crankangle and known 0 Spark Timing(deg) 328 flame radius can be found. Therefore, geometric 0 300 600 900 1200 0 300 600 900 1200 interaction between a flame and a combustion chamber BMEP (kPa) BMEP (kPa) with varied volume can be modeled with this (c) NO (d) Spark Timing methodology. 1.8 500 MAP: 0.34 bar MAP: 0.98 bar

Calibration of SIS 1.2 250

Cylinder pressure is used to calibrate the model 0.6 0 constants at various operating points because it captures PRP

S_laminar (m/s) S_laminar 9.26_Base global effects of turbulence and heat transfer losses. The 0 (kPa/deg.) dP/dtheta -250 results of SIS calibration are shown in Fig. 6. Pressure 300 360 420 480 540 240 300 360 420 480 Crankangle (deg) Crankangle (deg) diagram based on crankangle shows good agreement between SIS predictions and experimental results, as (e) Laminar Flame Speed (f) Pressure Rise Rate shown in Fig. 6 (a). In addition, cycle simulation Fig. 7 Effect of Engine Load at 1400 rpm with First predicted IMEP in reasonable agreement with actual test Spring data as shown in Fig. 6 (b). Therefore, SIS calibration demonstrates the robustness of this program to be useful The overall performance of the PRP engine with the for assessing the effect of spring set of PRP engine. first spring set is changed from the upper boundary to the 45 1100 1100 lower boundary, as load increases at 1400 rpm, which is exp IMEPg,ex p SIS IMEPg,s is shown in Fig. 7 (a), (b), and (c). In Fig. 7 (b), the BSFC BMEPexp 30 800 800 BMEPsis improvement of the PRP engine over a fixed CR engine operating at the lower PRP boundary is 27.2% at light 15 Pressure (bar)

IM EP (kPa) 500 500 BMEP (kPa) BMEP load (0.34 bar MAP). The NO increase of the PRP engine due to higher CR at low load is offset by retard of spark 0 200 200 120 240 360 480 600 timing, as shown in Fig. 7 (c). MBT timing of the PRP 20 40 60 80 100 engine is less sensitive to load compared to the Crankangle (deg) Load (%) conventional engines as shown in Fig. 7 (d). It is because (a) PRP Engine at 50 % load (b) Baseline spark timing advance due to lower engine load is offset Fig. 6 Calibration of SIS at 2000 rpm by higher CR and smaller residual gas fraction, which tend to retard the spark timing. Parametric Study Higher cylinder pressure during the early part of combustion, and smaller residual gas fraction due to 1. Engine Load Effects higher CR, yield the faster laminar flame speed of the PRP engine, as shown in Fig. 7 (e). Therefore, the PRP Conventional engines with fixed compression ratio engine has a shorter ignition delay compared to the equal to the upper and lower CR boundaries attainable conventional engine. The combustion noise of the PRP with the PRP engine are also simulated and included in

Presentation No. 81 Speaker Name: W. CHO 4/6 20056083 engine is expected to be reduced at WOT, because Fig. 8 (g) shows a relative NO increase of the PRP pressure rise rate is 50% smaller than a fixed CR engine engine with the first spring compared to a fixed CR operating at the lower PRP boundary at 1400 rpm, as engine operating at the lower PRP boundary (CR=9.26) at shown in Fig. 7 (f). various engine speeds and loads. As load is increased, the relative NO increase of the PRP engine is reduced, while 2. Engine Speed Effects NO increase of the PRP engine is independent of engine speed. The NO emissions of the PRP engine at high loads The results of an engine speed sweep carried-out is lower than those of the lower boundary case because with the first spring set are plotted in Fig. 8. The PRP the ECR of the PRP engine is lower at high loads due to engine achieves its minimum BSFC at mid engine speed, less stiffness of the first spring set. As illustrated in Fig. 8 and its value increases as engine speed is increased or (h), the BSFC improvement of the PRP engine is larger decreased, as shown Fig. 8 (a). Heat transfer loss at low with increasing engine speed, as well as with decreasing engine speed and rubbing friction at high engine speed engine load. Inertia force is proportional to engine contribute to the increase of BSFC, as plotted in Fig. 8 (b) speed squared and is exerted on the piston upwards and (c). PMEP shown in Fig. 8 (d) increases with engine during the combustion period. Therefore, inertia force has speed, because the pressure drop across the intake and an effect of increasing CR. In addition, its influence is exhaust valves, see Fig. 8 (e), increases as the flow rate greater at part load because the cylinder pressure force is across these restrictions increases. Exhaust temperature in relatively smaller at part load. Another reason is the Fig. 8 (f) is increased with engine speed due to higher induction tuning effect of the Hydra engine at high engine frequency of heat release and exhaust. Also, there is less speed. More air can be inducted at high engine speeds time for heat transfer from exhaust gas to port as engine due to tuning effect, resulting in more power. speed is increased. 3. Spring Preload Effects 300 23 The relationship between spring set constant and 290 21 preload used for this study are illustrated in Fig. 9 (a). 280 MAP: 0.98 bar The peak spring force is fixed as 33360 N because the MAP: 0.74 bar 19 270 exceeding values are unpractical, considering structural

BSFC (g/kW/hr) BSFC 17 issues. Therefore, the spring constant is adjusted Heat TransferHeat Loss (%) 260 1000 2000 3000 4000 5000 depending on the preload. 1000 2000 3000 4000 5000 Engine Speed (rpm) Fuel conversion efficiency is gradually increased Engine Speed (rpm) with the increase of the spring set preload and does not (a) BSFC (b) Heat Transfer Loss increase beyond a certain preload, as shown in Fig. 9 (b). 250 -5 This is because the spring set preload beyond this point is

230 -15 greater than the peak pressure force that the spring set is no more deflected and the ECR of the PRP engine 210 -25 maintains constant. Therefore, there is no more FMEP (kPa) FMEP 190 PMEP (kPa) -35 improvement of fuel conversion efficiency above this preload. As shown in Fig. 9 (c), NO concentrations 170 -45 1000 2000 3000 4000 5000 1000 2000 3000 4000 5000 follow the trend of fuel conversion efficiency at part Engine Speed (rpm) Engine Speed (rpm) loads, but they are less sensitive to the spring set preload at high load because the peak force of the spring set is (c) FMEP (d) PMEP constrained as constant. In Fig. 9 (d) the effect of the 150 1280 spring set preload on fuel consumption is illustrated at 1400 rpm 130 1255 2600 rpm. The BSFCgain in Fig. 9 (d) is defined as 3800 rpm following: 110 1230 BSFCPRP − BSFCbaseline BSFCgain = ×100 (%) (7) 90 1205 BSFC Pressure (kPa) baseline 1180 70 As shown in Fig. 9 (d), the maximum BSFCgain of the Exh. Temperature (K) Exh. Temperature 1000 2000 3000 4000 5000 0 200 400 600 PRP engine is obtained with a higher spring set preload Volume (cm^3) Engine Speed (rpm) occurs under part load conditions. It is because the higher (e) PV Diagram (f) Exhaust Gas Temperature ECR of the PRP engine induced by higher spring set

70 % 50 preload results in lower fuel consumption due to 1400 rpm % 2600 rpm increased thermal efficiency. As load is increased, the 40 30 3800 rpm fuel economy improvement of the PRP engine is decreased because of the decrease of the ECR of the PRP

10 10 engine. As the spring set preload is increased, the increase of BSFCgain of the PRP engine at part loads is smaller NO Increase ( of PRP BSFC Improvement BSFC ( of PRP -20 -10 than at high load cases. It is because the ECR of the PRP 0.2 0.4 0.6 0.8 1 0.2 0.4 0.6 0.8 1 engine at part load is initially increased as the spring set MA P ( bar ) MA P ( bar ) preload is increased. However, it reaches the highest CR (g) NO Emission (h) BSFC as preload is increased further, because of the relatively Fig. 8 Effect of Engine Speed lower cylinder pressure as shown in Fig. 9 (e). Even

Presentation No. 81 Speaker Name: W. CHO 5/6 20056083 though the maximum deflection is almost the same 2. Pouliot, H. N., Delameter, W. R., Robinson, C. W., regardless of preload at full load, the spring deflection "A Variable-Displacement Spark-Ignition Engine", duration of the PRP engine is smaller with the increase of SAE 770114, 1977 preload, resulting in more work, as shown in Fig 9 (f). 3. Kentfield, J.A.C., "Diesel Engines with Extended 40 Expansion Strokes", SAE 891866, 1989 43 4. Kutenev, V., Romanchev, Y., Zlenko, M., "Axial 30 40 Internal Combustion Engine - Practical Prospects for preload 20 the Future", SAE 940204, 1994 700 37 0.34 0.52

Load (kN) 8100 MAP 0.7 0.88 5. Larsen, G.J., "Research Engine for Evaluating the 10 15500 Nfuel (%) 34 22900 Effects of Variable Compression Ratio (VCR) and/or 30300 0 31 Variable Valve Timing (VVT)", SAE 910053, 1991 0 0.09 0.18 0.27 0.36 0 11000 22000 33000 6. Drangel, H, Olofsson, E., Reinmann, R., "The Deflection (cm) Preload (N) Variable Compression (SVC) and the Combustion (a) Load vs. Deflection (b) Fuel Conversion Efficiency Control (SCC) ~ Two Ways to Improve", SAE 2002- 28 01-0996, 2002 2500 preload 700 7. Adams, W. H., Hinrichs, H. G., Pischinger, F. F., 21 8100 2000 15500 Adamis, P., Schumacher, V., Walzer, P., "Analysis 22900 of the Combustion Process of a Spark Ignition 1500 14 30300

NO (ppm) NO Engine with a Variable Compression Ratio", SAE 1000 BSFC (%) gain 7 870610, 1987 8. Wong, V. W., Stewart, M., Lundholm, G., Hoglund, 500 0 0 11000 22000 33000 0.2 0.4 0.6 0.8 1 A., "Increased Power Density via Variable preload (N) Load (MAP) Compression/Displacement And Turbocharging

(c) NO (d) BSFCgain Using The Alvar-Cycle Engine", SAE 981027, 1998 14 3 9. Grundy, J. R., Kiley, L. R., Brevick, E. A., "AVCR 700 1360-2 High Specific Output- Variable Compression 13.5 11800 Ratio Diesel Engine", SAE 760051, 1976 2 30300 R 13 10. Ashley, C., "Variable Compression Pistons", SAE EC 1 901539, 1990 12.5 Deflection (mm) 11. Wirbeleit, F.G., Binder, K., Gwinner, D.,

12 0 "Development of Pistons with Variable Compression 0 8000 16000 24000 32000 330 360 390 420 Height for Increasing Efficiency and Specific Power Preload (N) Crankangle (deg) Output of Combustion Engines", SAE 900229, 1990 (e) ECR (f) Deflection 12. Rychter, T. J., Teodorczyk, A., "VR/LE Engine with Fig. 9 Effect of Preload on Engine Performance a Variable R/L during a Single Cycle", SAE 850206, 1985 Conclusions 13. Kajiwara, K., “A Variable-Radius/Length Engine”, SAE 920453, 1992 • At light load enough for the spring not to be 14. SAE HS 1582 Part 5, “Design and Manufacture of deflected, the best improvement of the PRP engine was Coned Disk Springs (Belleville Springs) and Spring observed because the PRP engine performance is almost Washers” same as an engine with fixed compression ratio equal to 15. Cho, W., “A Study on Pressure Reactive Piston for the upper CR boundaries Spark Ignition Engines”, Ph.D. Thesis, 2004 • The upper piston motion of the PRP engine at 16. Poulos, S. G., Heywood, J. B., "The Effect of high load caused a much smaller rise in cylinder pressure Chamber Geometry on Spark-Ignition Engine compared to an engine of fixed compression ratio equal to Combustion", SAE 830334, 1983 the lower CR boundary of the PRP engine. 17. Assanis, D. N, Filipi, Z. S., "Quasi-Dimensional • The BSFC improvement of the PRP engine Computer Simulation of the Turbocharged Spark- compared to a conventional engine increased with engine Ignition Engine and its Use for 2- and 4- Valve speed due to the inertia force on the piston crown with an Engine Matching Studies", SAE 910075, 1991 effect of increasing CR. 18. Tabaczynski, R. J., Ferguson, C. R., Radhakrishnan, • The fuel conversion efficiency was gradually K. “A Turbulent Entrainment Model for Spark- increased with spring preload until the spring could not be Ignition Engine Combustion” SAE770647, 1977 compressed any more because the preload beyond this 19. Blizard, N. C., Keck, J. C., "Experimental and point is greater than the peak pressure force. Theoretical Investigation of Turbulent Burning Model for Internal Combustion Engines", SAE References 740191, 1974

1. Schwaderlapp, M., Habermann, K., Yapici, K.I., “Variable Compression Ratio- A Design Solution for Fuel Economy Concepts”, SAE 2002-01-1103, 2002

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