THE UNIVERSITY OF MICHIGAN INDUSTRY PROGRAM OF THE COLLEGE OF ENGINEERING THERMAL LOADING AND/WALL TEMPERATURES AS FUNCTIONS OF PERFORMANCE OF TURBOCHARGED COMPRESSION: IGNITION ENGINES Naeim Abdou Henein This dissertation was submitted in partial fulfillment of the requirements for the degree of Doctor of Philosophy in the University of Michigan, 1957. July, 1957 IP-229

ACKNOWLEDGEMENTS It is a pleasure to acknowledge all those who have offered assistance and encouragement during the course of this investigation. The author is particularly indebted to Professor E. T. Vincent, the Chairman of the Doctoral Committee, for his continued active interest and suggestions which were a major factor in the completion of this work. Professor J. A. Bolt and W. H. Graves were important in helping with the experimental portions of the work. The author wishes to thank Professor R. B. Morrison and Professor R. A. Wolfe for their cooperation. Help in preparation of this report by the Industry Program of the University of Michigan is gratefully acknowledged. The author also wishes to express his appreciation to the men of the Automotive Laboratory shop who, despite the pressure of other duties in the new building, were of help in building the experimental equipment. i

TABLE OF CONTENTS Page ACKNOWLEDGEMENTS.................................................. i LIST OF TABLES.................................................... iii LIST OF FIGURES................................................... NOMENCLATURE...................................................... ix I. INTRODUCTION............................................... 1 II. THEORY OF HEAT TRANSFER IN THE ENGINE-CYLINDER............ 3 III. EXPERIMENTAL SET-UP........................................ 15 IV. EXPERIMENTAL PROCEDURE..................................... 37 V. EXPERIMENTAL RESULTS....................................... VI. DISCUSSION OF THE EXPERIMENTAL RESULTS.................... 6 VII. HEAT TRANSFER ANALYSIS OF THE EXPERIMENTAL RESULTS......... 66 VIII. INVESTIGATION ON THE EFFECT OF AFTERCOOLING................ 76 IX. CONCLUSIONS AND RECOMMENDATIONS............................ 92 X. APPENDICES: A. Sample Calculations.................................... 95 B. Calculation of Constant "C" of Equation [2.11]......... 1 C. Calculation of Combustion Chamber Wall Temperature..... 113 D. Calculation of Intensity of Thermal Load on the Combustion Chamber Walls............................... 116 E. Calculation of Mean Wall-Temperatures.................. 131 F. Scavenging Air Flow-Rate............................... 140 G. Sample Calculations for the Effect of Aftercooling..... 143 H. CorrectiDn Factor for the B.M.E.P...................... 155 I. Engine Specifications................................... 157 BIBLIOGRAPHY...................................................... 158 ii

LIST OF TABLES Table Page I. Heat Transfer Analysis for the Cycle, (Run #95).......o.. 106 II. Heat Transfer Analysis for the Cycle, (Run #95).......... 107 IIIo Engine Test Results, for: Pm: 30" Hg, Tm: 80~F & N: 800 R.P.M........ 117 IV. Engine Test Results, for: Pm: 33" Hg, Tm: 80~F & N: 800 R.P.M......o 118 V. Engine Test Results, for: Pm: 36" Hg, Tm: 80~F & N: 800 R.P.M........ 119 VI. Engine Test Results, for: Pm: 39" Hg, Tm: 800F & N: 800 R.P.M....... 120 VII. Engine Test Results, for: Pm: 42" Hg, Tm: 80~F & N: 800 R.P.M........ 121 VIIIo Engine Test Results, for: Pm: 45" Hg, Tm: 80~F & N: 800 R.PoM........ 122 IX. Engine Test Rcgsults, for: Pm: 36" Hg, Tm: 80~F & N: 1200 R.P.M........ 123 X. Engine Test Results, for: Pm: 36" Hg, Tm: 140~F & N: 1200 R.P.M....... 124 XIo Engine Test Results, for: Pm: 36" Hg, Tm: 200~F & N: 1200 R.P.Moo....o. 125 XIIo Engine Test Results, for: Pm: 36" Hg, Tm: 80~F & N: Variable.......... 126 XIIIo Engine Test Results, for: Extra runs referred to in calculations...o..... 127 XIVo Calculation of Constant "C" in Equation [2.11]............ o 28 XVo Calculation of Constant "C" in Equation [2.11].....,..... 129 XVIo Mean Coefficients of Heat Transfer and Mean Effective Temperatureso o.0.O. o o.O...oo..OOo 134 XVIIo Mean Coefficients of Heat Transfer and Mean Effective Temperatures........................ o 135 XVIIIo Combustion Chamber Wall Temperatures: Calculated and Measured..0..0.0....0..0...0.... 136 iii

LIST OF TABLES (CONT'D) Table Page XIX. Combustion Chamber Wall Temperature: Calculated and Measured........................ 137 XX. Calculated Thermal Loadings and Measured Heat Losses to Cooling Water......................... 138 XXI. Calculated Thermal Loading and Measured Heat Losses to Cooling Water......................... 139 XXII. Piston-crown Temperatures and Intensity of Thermal Loads, for e 0...................... 145 XXIII. Piston-crown Temperatures and Intensity of Thermal Loads, for c = 50...................... 146 XXIV. Piston-crown Temperatures and Intensity of Thermal Loads, for e = 100..................... 147 iv

LIST OF FIGURES Figure Page 1 Gas and Cylinder Wall Temperatures...................... 7 2 Areas of Heat Transfer.................................. 7 3 Graphical Construction for the Imaginary Walls.......... 10 4 Graphical Construction and Temperature Fields for the Piston................................ 12 5 General Layout of Experimental Set-Up................... 16 6 Side View Showing the Air Flow Meter and the Intake System................................. 17 7 Instrumentation......................................... 18 8 General View of the Oscilloscope and Amplifiers......... 19 9 Diagram of Intake System................................ 20 10 Cooling Water System.................................... 23 11 Automatic Fuel Weighing and Revolutions Counting Devices 25 12 The Degree Marking Unit................................. 27 13 Layout of Electrical Circuits for Pressure and Temperature Recording.......................... 2 14 "Tyni-Couple" Assembly.................................. 30 15 The Basic "Tyni-Couple" Unit............................ 31 16 Sectional Plan of the Cylinder Head..................... 32 17 Sectional Elevation of the Cylinder Head............... 33 18 Calibration Curve for the Combustion Chamber Thermocouple............................... 34 19 Thermocouples Positions on the Cylinder-Liner Walls..... 35 20 Gas Pressure During the Cycle........................... 41 21 Combustion Chamber Surface Transient-Temperature........ 42 v

LIST OF FIGURES (CONT'D) Figure Page 22 Combustion Chamber Surface Transient-Temperature........ 43 23 Pressure and Temperature Calibration Traces............. 44 24 Effect of F/A Ratio on B.M.E.P. for Various Manifold Pressures.......................... 46 25 Effect of F/A Ratio on I.M.E.P. for Various Manifold Pressures............................ 47 26 E.M.E.P. x 14.7 x Tm vs F Ratio for Various L^ Pm 7~1 A Manifold Pressures............................ 48 27 B.S.F.C. vs E Ratio for Various Manifold Pressures... 49 A 28 Mechanical Efficiency vs F Ratio for Various Manifold Pressures.A......................... 50 29 Heat Losses to Cooling Water vs F Ratio for Various Manifold Pressures... A..................... 51 30 Heat Losses to Cooling Water vs I.M.E.P. for Various Manifold Pressures............................ 52 31 Exhaust Gas Temperature vs F for Various Manifold Pressures.................................... 53 32 I.Th. vs I.M.E.P. for Various Manifold Pressures..... 54 33 B.M.E.P. vs F Ratio for Various Intake Air Temps..... 55 34 I.M.E.P. vs E Ratio for Various Intake Air Temps..... 56 A 35 [.M.E.P. x Tm] vs F Ratio for Various Intake Air Temperatures.................................. 57 36 B.S.F.C. vs F Ratio for Various Intake Air Temps..... 58 A 37 Heat Losses to Cooling Water vs I.M.E.P. for Various Air Temperatures............................. 59 38 Exhaust Gas Temperature vs E Ratio for Various Intake Air Temperatures............................. 60 vi

LIST OF FIGURES (CONT'D) Figure Page 39 aM. vs I.M.E.P............................. 68 40 _. _ vs I.M.E.Pe................................ 69 5Us TTm 41 Check on TM.E and xM................................... 70 42 Diagram of assumed Turbocharged C.I. Engine.......... 77 45 Performance with 100% aftercooler Effectiveness......... 80 44 Performance with 50% aftercooler Effectiveness...... 81 45 Performance without aftercooling....................... 82 46 Effect of Aftercooler Effectiveness for Pm = 45" Hg..... 85 47 Effect of Aftercooler Effectiveness for Pm = 45" Hg..... 86 48 Effect of Aftercooling on Intensity of Thermal Loading and Piston Maximum Temperature........ 89 49 Effect of Aftercooling on Power Output................. 90 50 Effect of Aftercooling on Reduction of Thermal Loading and Piston Maximum Temperature................ 91 51 (P-V) Diagram for Run Number 95......................... 101 52 Coefficient of Heat Transfer from Gas to Walls for the Cycle of Run No. 95........................... 108 53 (CagTg) for the Cycle of Run No. 5.................... 109 54 Surface Area Variation vs Crank Angles................ 110 55 Calculation of the Average Wall Temperature for Run No. 95........................................ 112 56 Calculation of Twg and Tpg.............................. 115 57 Calculation of Constant "C" in Equation [2.11].......... 130 58 Calculation of Cylinder Walls Mean Temperature for Run No. 95.................................... 133 vii

LIST OF FIGURES (CONT'D) Figure Page 59 Scavenging Air Flow Rate Diagram......................... 42 60 Tp.g. vs I.M.E.P. for Different Manifold Pressures = oo %....................................... 148 61 q. vs I.M.E.P. for Different Manifold Pressures = oo...................... o................ 149 62 Tp.g. vs I.M.E.P. for Different Manifold Pressures, C = 50%........................................ 150 63 q. vs I.M.E.P. for Different Manifold Pressures, = 50o%....................................... 151 64 Tp.g. vs I.M.E.P. for Different Manifold Pressures, E -= 0%......................0.................. 152 65 q. vs I.M.E.P. for Different Manifold Pressures, e = 0%......................................... 153 66 q. vs I.M.E.P. for Different Effectiveness.............. 154 67 Correction factor for the B.M.E.P...... 156 viii

NOMENCLATURE a = Area factor for the coolant side of wall = A c Ac am = Area factor for the mean area of wall = Am Aa = Area of piston-crown on the crank-case side (sq. ins or sq. ft.) Abore = Area of cylinder bore. (sq. ins. or sq. ft.) A = Area of wall on the coolant side. (sq. ins. or sq. ft.) Ac.ch. = Area of combustion chamber. (sq. ins. or sq. ft.) Aexh. man.= Area of the exhaust manifold enclosed in the cylinder head (sq. ins. or sq. ft.) Ag == Area of wall on the gas side. (sq. ins. or sq. ft.) Aint. man.= Area of the intake manifold enclosed in the cylinder head (sq. ins. or sq. ft.) Am = Mean area of wall. (sq. ins. or sq. ft.) AM = Mean effective area of heat transfer. (sq. ins. or sq. ft.) P.T. = Area of piston top. (sq. ins. or sq. ft.) c = Specific heat. B.T.U. lb. ~F C and C1 = Constants. B.T.U. c = Specific heat at constant pressure. lb. F D = Diameter ins. or ft. E = Area multiplier for thermal expansion. lb s. G = Flow rate lbs. hr. sq. ft. hi = Enthalpy of air at compressor inlet B.T.U. lb. h2 = Enthalpy of air at compressor outlet B.T.U. h2' = Enthalpy of air after isentropic compression B.T.U. lb. ha = Enthalpy of air B.T.U. Ib. hexh. = Enthalpy of exhaust gases B.T.U. lb. ix

NOMENCLATURE (CONT'D) B.T.U. hmix = Enthalpy of mixture of air and exhaust gases. lbU lb. B. T. U. k = Thermal conductivity of water. B.T.U. hr. sq. ft. (-I) ft B. T. U. k, = Thermal conductivity of the wall metal. B.T.U. "~~~~~~~~w ~~hr. sq. ft. (~F ft K = Flow coefficient. m,n = Exponents. N = Revolutions per minute. NtF = Revolutions of crank shaft for the time period tF. P = Air pressure at compressor inlet - r ins. Hg 1 —~~~~~~~~~ ~sq. in. Pa = Air pressure at inlet tap of orifice meter lbs or ins. Hg. sq. in. Pb = Barometric pressure bs or ins. Hg. sq. in. Pm = Manifold Air pressure lbs or ins. Hg. Pr, Pr = Relative Air pressures lbs or ins. Hg. 1 2 sq. in. Q = Quantity of heat transfer B.T.U. hr. q = Intensity of thermal load B.T.U. hr. sq. ft. rm = Mean radius R = Universal gas constant ft. lb. lb. OR Re = Reynold's number S = Mean piston speed ft. sec. t = Time tF = Average time for consumption of 0.1 lb of fuel --- minutes. Ta = Air temperature at orifice meter ~F or ~R T1 = Air temperature at compressor inlet, or atmospheric ~F or ~R T2 = Air temperature at compressor outlet, or atmospheric ~F or ~R x

NOMENCLATURE (CONT'D) Tc = Cooling water temperature at inlet to barrel ~F or ~R T = Cooling water temperature at inlet to cylinder head ~F or ~R T - = Cooling water temperature at exit from cylinder head ~F or ~R c3 TE = Mean effective temperature over the suction stroke ~F or ~R Texh. = Exhaust gas temperature ~F or ~R Tg Gas temperature ~F or ~R T, a Liner wall outside surface temperature at T.D.C. ~F T = Liner wall outside surface temperature between T.D.C. and 2 B.D.C. ~F To = Liner wall outside surface temperature at B.D.C. ~F Tm = Air temperature in intake manifold. ~F or ~R TM.E. = Gas mean effective temperature over the whole cycle. ~F or ~R Tmix. Temperature of mixture of air and clearance exhaust gases ~F or ~R To = Oil temperature in crank case. ~F or ~R Tp.a. = Temperature of piston-crown on the crank-case side. ~F or ~R Tp.g. = Temperature of piston-crown on the gas side. ~F or ~R Ts = Average temperature of the engine outside wall. ~F or ~R Tw.c. = Wall temperature on the coolant side. ~F or ~R Tw.ex. = Wall temperature of the exhaust manifold (coolant side)~F or ~R Tw.g. = Wall temperature on the gas side ~F or ~R U = Overall coefficient of heat transfer B.TU, hr. sq. ft. ~F v = Velocity of flow ft. sec. V = Volume cu. ins. or cu. ft. Vs = Engine swept volume cu. ins. Ua = Weight of air per cycle xi

Wexh = Weight of residual exhaust gases per cycle Wa = Air flow rate lbs a hr WB - Brake load lbs. lbs WC = Cooling water flow rate -l WF = Friction load lbs. W = Indicated load lbs. x = Wall thickness ins. x = Piston-crown thickness ins. x + kw Y = Constant = a " aa z = Constant = amX + a kw c aT^ 0a_ = Coefficient of heat transfer from wall to air B.T.U. hr. sq. ft. ~F Qac = Coefficient of heat transfer from wall to coolant B.T.U. hr. sq. ft. ~F a = Coefficient of heat transfer from gas to wall B.T.U. hr. sq. ft. ~F aM = Mean coefficient of heat transfer over the whole cycle B.T.U. hr. sq. ft. ~F Pa = Density of air lb. cu. ft. Pc = Density of coolant lb. cu. ft. ~e = Aftercooler effectiveness = T2Tm T2-T1 I= Dynamic viscosity flb ft. hr. rcompL = Compressor Isentropic efficiency B.Th. = Brake thermal efficiency I.Th. = Indicated thermal efficiency m == Mechanical efficiency xii

NOMENCLATURE (CONT'D) List of Abbreviations and Definitions: B.D.C. = Bottom dead center. T.D.C. = Top dead center. F A = Fuel air ratio. lbs. B.M.E.P. = Brake mean effective pressure. sq. sq. in. I.M.E.P. = Indicated mean effective pressure. lbs sq. in. B.S.F.C. = Brake specific fuel consumption lbs. B.H.P. hr. I.S.F.C. = Indicated specific fuel consumption lbs. I.H.P. hr. B.H.P. = Brake horse power. I.H.P. = Indicated horse power. H.V. = Heating value of fuel B.T.U. lb. C.I. Engine: Compression ignition engine. Thermal loading equals the direct heat transfer from the gas to the enclosing walls per hour. xiii

I. INTRODUCTION The purpose of this investigation was to study the performance of the turbocharged compression ignition engines with emphasis on the fuel economy and the heat problem. The main idea of turbocharging is to compress air into the engine cylinder to increase its ability to burn more fuel; thus increasing its power output. The energy in the exhaust gases is used to drive a turbine which, in turn, drives the compressor. Due to compression, the air temperature rises to values higher than atmospheric. In some turbocharged units an after-cooler cools the air after it leaves the compressor to get more power from the engine. Besides increasing the engine power output turbocharging improves its efficiency; however, higher gas pressures and temperatures are reached in the cylinder. These tend to increase the mechanical loads on the engine parts and the thermal loads on the cooling system while the temperatures of the engine parts go up. These temperatures are a major factor in the engine operation because: a. High local temperatures in some hot spots in the combustion chamber or piston-top, causing excessive heat-stresses, might end with cracking. b. The temperatures of the piston and cylinder walls might get high enough to cause rapid evaporation of the piston lubricating oil film, thus injuring the piston and cylinder surfaces. c. The piston ring-temperatures might get high enough to cause them to stick or loose their springy action. -1

-2The above conditions pointed out the need to study the factors affecting the thermal loads and the wall temperatures, as well as investigating the way engine performance and fuel economy are improved. For this purpose a new single cylinder C.I. Engine at the Automotive Laboratory was chosen to be supercharged. Since applying a turbocharger to this engine was impossible; it was decided to investigate, individually, the effect of the different variables on the engine operation. Air from a compressed-air line was delivered to the engine, while an electric heater served to heat the air before it entry into the intake manifold. Besides the measurements necessary to study the engine performance; it was of interest to measure the transient surface temperature of the combustion chamber. A thermocouple manufactured to measure the bore temperature of gun barrels was used for this purpose. Pictures for the surface transient-temperature and the gas pressure during the cycle were taken for each run. As a result of this study some formulae were derived to calculate the power output, wall inside surface temperature, and intensity of thermal loads of the turbocharged engine. The combustion chamber temperatures and the heat losses calculated from the derived formulae were compared with those measured. Applying these formulae an investigation was made of the effect of aftercooling on the turbocharged engine power output, thermal loading and wall temperatures.

II. THEORY OF HEAT-TRANSFER IN THE ENGINE-CYLINDER The heat transfer in the engine-cylinder occurs in three steps: A. Heat transfer from the gases to the walls. B. Heat transfer through the walls. C. Heat transfer from the walls to the cooling medium. In this chapter, the process of heat transfer in each of these steps will be discussed separately. The final equation for the overall heat transfer will be used to find the combustion chamber wall temperature. Other equations will be given for the maximum temperature )f the piston top, the intensity jf thermal loading on the combustion chamber walls, and the thermal loading on the cylinder walls. A. Heat-Transfer from the Gases to the Wall: The rate of heat transfer from a gas to a surrounding surface is, in general, a function of many factors such as: the surface area, the temperature difference between the gas and the surface, and a coefficient of heat transfer. For any internal combustion engine, these factors change from instant to instant throughout the cycle. a. Surface Area Variations: The surface exposed to the gases in the cylinder varies from a minimum at top dead center to a maximum at bottom dead center. When the piston is at the top dead center, the surface exposed to the gases cinsists of that of the c mbustion chamber plus the piston top. As the piston moves toward the bottom dead center, the cylinder bore is exposed to the gases. -5

-4The surface area variation of the engine, used in tests, is given in figures (2) and (54). The area during the intake and exhaust strokes includes the portions of the intake and exhaust manifolds enclosed in the cylinder head respectively. These manifold walls transfer heat between the gases and the cooling water during the corresponding strokes. b. Gas Temperature Variations: The temperature of the gases in the cylinder varies appreciably during the different strokes. At the beginning of the intake stroke, the temperature is that of the clearance gases, but falls rapidly as the fresh air is brought in. It rises during the compression stroke, reaches its maximum at the end of the combustion process, then decreases with expansion and drops rapidly after the exhaust valve opens. There is a small drop in the temperature during the exhaust stroke. c. Wall Temperature Variations: The temperature Jf the inside surface of the cylinder wall fluctuates during the cycle following the variation of the gas temperature. The wall temperature goes to its maximum at the end of the combustion process, then drops continually during the exhaust and intake strokes, and reaches its minimum during the early part of thi compression stroke. Figure (21) shows a picture of the wall temperature variations for the whole cycle of the engine during run No. 95,

whose conditijns are given in Appendix A. d. Coefficient of Heat Transfer Variations: The change in the coefficient of heat transfer, during the cycle, follows the variations in the gas temperature and pressure. At any instant of the cycle this coefficient is a function of the instantaneous pressure and temperature of the gas. This function was given by G. Eichelberg(9): C^ = 2.1 ^/S J^ -K. cal. g g ~PThr. m2. ~C. where S is in meters per sec. P in atmospheres. Tg in degrees Kelvin. This equation was put in B.T.U.s as follows: = o.0564 s s B.T.U. ~~~g ~~ ~hr. sq. ft. ~F. [2.1] where S is in feet per sec. P in lbs. per sq. in. T T in degrees Rankine. G g Equation of Heat Transfer: GAS I, COOLING The heat transfer from the gas to SIDE | ATR Z WATER the cylinder walls is given by, t SIDE ti = c.A.(T -.T )dt W.G. TW.C. g g w.g. T Writing this equation for each stroke: * References are given in Bibliography.

-6(i) Intake Stroke: t Q = [A-.h. + Ap.T. + AInt.Man.]j g(Tg-Tw.g.)dt tl + bore g (Tg-Tw.g.)dt [2.2 0 (ii) Compression Stroke: Q = [A.h A + A.T. g (Tg-.g) re Tg g g)dt t tl [2.3] 1 (iii) Expansion Stroke: Q=[A,, + A.T~. ll~" it3 Q = [A.h. + Ap.T.]J g (Tg-T.g.)dt + Abore ag (TgTw.g.)dt t t [2.4] (iv) Exhaust Stroke: t4 = [A h. + Ap.T. + AExh.Man.] f (Tgw.g. )dt t4 t3 + Abore g (Tg-Tw.g)dt t3 [2.5] Since heat transferred during the cycle = the sum of heat transferred during the four strokes, then by adding equations [2.2], [2.3], [2.4] and [2.5] we get: Equation of Heat Transfer to the Combustion Chamber Walls: t 4 Qc.ch. = Ac.ch. g (Tg-Tw.g.)dt [2.6] and the mean heat transferred t4 -. ch. ag (T -T )dt g w.g. 0 In this integral the wall temperature T can be considered as constant, and the equation becomes:

-7Gas Temperature Cylinder Wall Temperature T.D.C. B.D.C. T.D.C. B.D.C. T.D.C. 0 t t2 t3 t4 Intake Compression Expansion Exhaust Figure 1. Gas and Cylinder Wall Temperatures. / Ie Manifold haust Manifold Cylinder B/re Combustion Chamber+Piston Top Intake Compression Expansion Exhaust Figure 2. Area of Heat Transfer.

-8Qc.ch. = Acch. [ (gTg)M - aM Tw.g.] = Ac.ch. cM [TM.E. - T.g.] [2.7] where TM E = the mean effective temperature = (cgTg)M [2.8] CXM = the mean coefficient of heat transfer from the gases to the walls. B. Heat Transfer Through the Walls: The net heat given by the gases to the surrounding surface is transferred by conduction through the walls. The equation of heat transfer by conduction is, Q = A kw (T.g.- Tw.c) [2.y] and T,. can be considered constant as in equation [2.6]. w.g. C. Heat Transfer to the Cooling Medium: The heat conducted through the walls is transferred to the cooling medium, mainly, by convection. The equation of heat transfer by convection is, Q = c Ac (Tw.c. - Tc) [2.10] where Xc, is a function of: the rates of flow of the cooling medium; its physical properties, and the mechanism of cooling. For an engine

_0_ cooled with water, Czc can be calculated by using the Nusselt's equation: acD (vDp)m ( c)n K K2 For any engine: D, is constant; and vp, is proportional to the rate of flow of the cooling water. If the cooling water temperature is thermostatically controlled; K will be constant, and Nusselt's equation will take the form: C= Wc (c) [2.11] (51) The values of the exponents m and n were given by B. Pinkel m = 0.6, n = 0.4 and C is a constant of each engine. Equation for the Combustion Chamber Wall Temperature: Equating for Q from equations [2.7], [2.9] and [2.10] we get: Ag. M [TM.E. - Tw.g.] -= kw [Tw.g. - T w.cJ A= A c [Tw.c. - Tc ] These equations can be arranged in the following form, kw A x A kw = ( )x x tM [TM.E.-Tw. g- ] (Tw.g.-Tw.c.) AC a [Tw. c-Tc [2.12] Equation [2.12] is represented graphically in figure (3), by the slope of the straight line between TME. and T. M. E. c

-10EFFECTIVE WALL THICKNESS Am G. Tw ( GAS SIDE. c WATER SIDE IMAGINARY WALL | /l IMAGINARY WALL --- aM —----'7 Oc Ac Figure 3 Graphical Construction for the Imaginary Walls Using the following area factors: am = Am A and ac = -g Ac the following equation can be derived from figure (3): TM.E._- - T Tw' - Tc [2.15] (aw) + a k a x + a kw ] m C c m c Let Z = [am x + ac w] T -T z [W,. _c1 _ zk T.M.E -Tc] Z + TM and the equation of combustion-chamber-wall inside-surface temperature will be:

-11Tw. TM.E. + T [ M [2.14]. Z + kw k + Z M l Intensity of the Thermal Loading on the Combustion Chamber Walls: This is given by the amount of heat transferred to the walls per unit area per unit time. q = U (TM.E.- T) [2.15] where U is the overall coefficient of heat transfer between the gases and the cooling medium, and can be calculated from equations [2.7], [2.9] and [2.10] as, 1 1 x 1 [2.16] - + am'- ~ - m kw M kw C Thermal Load on the Cylinder Walls: This is given by the heat transferred per unit time to the engine walls surrounding the gases. Q = AMU [TM.E.- Tc] [2.17] where AM is the mean effective area for heat transfer, and is given by: t4 f ag A(Tg-TW g) dt j ag (Tg-Tw.g.) dt it and U is the same as that in equation [2.16].

-12Lines of Constant I + i FTemperature Lines of CHeat Flow Figure (4,a) Temperature Field in the Piston-Crown. T.E. Ag Aa.a, TTp Ta kw kw * "M''// /. "aI Figure (4,b) Graphical Construction of the Imaginary Walls for the Piston Crown.

-13Equation for Piston Top Maximum Temperature: Part of the heat transferred from the hot combustion gases to the piston top, goes to the mixture of the hot air and oil vapor on the other side of the crown. The rest goes to the cylinder liner, through the piston rings and the lubricating oil film. This is illustrated in figure (4,a) by the temperature fields calculated for similar types of pistons.(9) The maximum piston temperature occurs at the center of the piston-crown, as it is the point fartherest from the cylinder liner. Applying equation [2.13] to a small area at the center, with aa in place of ac' we get: TM - T T - T M.E. a p.g. a kw ++ [x + 1 "M P [XpAm pa Aa PAm a Aa ] [2.19] A A To get A- and A_, the temperature field in the crown has to be Am Aa calculated. Since the surface temperature is changing for each cycle, these calculations will turn out to be very complicated. Yet for the purpose of comparison between the piston temperatures at the different turbocharging conditions, the following assumptions will be made: Ag A 1 and a =1 Then equation [2.19] becomes: TME. Ta Tpg. -Ta kw kw kw [2.20] (&^) + [Xp+ ^] [Xp + a, ]

where xp and kw are constants for any engine, and aa is constant for constant crank oil temperatures kw Let Y [ + xp Ua T M.E. =a] T l-Ta -k + y Y CAnd And the equation for the maximum temperature of the piston-crown will be: k kw Y _________M T T [ i +T [ p =g TM.E. Y+kw [w + Y] [2.21] aM 4M The wall temperatures of other parts of the cylinder can be evaluated by a similar analysis if their cooling mechanism is known. For the above analysis the gas mean effective temperature and the gas mean coefficient of heat transfer were evaluated, from the cycle analysis, as functions of the following performance conditions: a - The indicated mean effective pressure. b - The intake manifold pressure. c - The intake manifold temperature. d - The mean piston speed, The thermal loading and wall temperatures were then calculated and compared with the experimental results, as shown in the following chapters,

III. EXPERIMENTAL SETUP A single-cylinder, 4-stroke cycle, liquid-cooled Nordberg Diesel Engine was used in the experimental work. The cylinder had a bore of 4 1/2 inches, a stroke of 5 1/4 inches, and a compression ratio of 14.5, (Other engine specifications are given in Appendix I.). The general test setup is shown in figures (5) (, () (7) and (8). The various systems are described in the following paragraphs. A. Air Intake System: The air intake system-is shown schematically in figure (9). Air under pressure entered the system through a filter to eliminate entrained solids, and then passed through an automatic pressureregulating valve before entering the air flowmeter. This consisted of an A.S.M.E. sharp-edged orifice in a flange mounting, figure (6). The orifice disc was of stainless steel and the Drifice was 0.7018 inch diameter. The conduits on both sides of the orifice disc were of stainless steel tubing of 2.0052 inches inside diameter. The length of the straight tube ahead of the orifice was 36 inches and was used to secure a fairly uniform velocity profile in the approaching stream. Air leaving the flow meter passed through an electric-heater to be heated to the required temperature. The heater shown in Figures (5), (6) and (8) has a capacity of 3000 watts. It had 3 coils connected in parallel through a switch, which allowed reduced power outputs of 1500 or 750 watts. A copper-wool chamber and a surge tank were placed between the heater and the engine intake manifold, to reduce pulsation to a -15

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o ~ l G L A. Main Air Supply Valve F. Air Line to Pressure Pick Up I. Air Flowrneter and Manometers B3. Main? Air Pr~essure Gag;e F. Air Su~pply Valve J. Electric Heater C. Con~densate Blood Valve G. Air Sulpply Control Valve K. Copper-Wool Chamber D. Air Filter HI. Air Pressure Gage L. Surge Tank M. Engine Fig'ure 9. DIAGRAM OF~ INTAKE SYSTEM. ^A ~H D G A. Main Air Supply Valve E. Air Line to Pressure Pick Up I. Air Flowmaeter adMnmtr B. Main1- Air Pressure Ga,-Ie F. Air Su~pplyr Valve J. Electric Heate C. Co-n~densatle Bloed Valve G. Air Supply Control Valve K. Copper-Wool Chme D. Air Filter II. Air Pressure Gar-,e L. Surge Tank M. Engine Fig~ure 9. DIAGRAM OF, INTAKE SYSTEM.

-21minimum. The copper wool chamber was 4 inches in diameter and 4 inches long. The surge tank was 15 inches in diameter and 19 inches long; with a volume 40.5 times the engine swept volume. The air pressure was measured before the flowmeter and in the surge tank by a mercury manometer. The pressure drop across the orifice was measured from two flange taps connected to a water manometer. The air temperature was measured by iron- constantan thermocouples, placed at the flowmeter and in the intake manifold of the engine. B. Exhaust System: For most of the runs a flexible exhaust pipe was connected between the engine exhaust manifold and the main exhaust underground conduit. For runs requiring back pressures, a surge tank was connected between the engine and the main conduit. The discharge pipe from the surge tank had a 2 inch gate valve to throttle the exhaust gases flow; thus building the required gas pressure. The pressure in the surge tank was measured by a mercury manometer. The gas temperature was measured by a chromel-alumel thermocouple placed in the exhaust manifold just after it leaves the cylinder head block. C. Cooling System: A cooling water closed system, equipped with a heat exchanger, was used to cool the engine. It is shown schematically in figure (10). It consisted of a pump driven by the engine, a heat exchanger, an automatic thermostatic control with a manual temperature control, and a water flow meter.

-22The manual temperature control was set to keep the water temperature at about 160~F, as it entered the engine. A 2 inch A.S.M.E. sharp edged orifice, similar to that used in the air inlet system, was used to measure the water flow rate. The temperature of the water was measured before entering and after leaving the engine, and in three passages leading from the cylinder barrel to the cylinder head to separate barrel and head losses.

er flow Glass Tube J So t -'-' - i — Engine G' Heat Exchanger Line from water maine Calibrated Orifice 0 t Thermostatic Element Figure 10. Cooling Water System.

-24Fuel Weighing System: The fuel weighing system measured the time required for the engine to consume definite weights of fuel. The system, figure (11), consisted of a scale which was equipped with a mercoid contactor. This operated electrical circuits which started and stopped an elapsedtime electric clock, and a revolution counter. The scale, clock, and counter measured the time required for the engine to consume a predetermined weight of fuel, and the number of revolutions of the crank shaft during the same period of time. This system had been used at the Automotive Laboratory, University of Michigan. Power Absorbing Unit: A cradle type direct current dynamometer, was used to start the engine and absorb the power developed. When operated as a generator it measured the power input and turned it into electrical energy, which was converted into heat in loading resistance grids. The power was measured by the pull exerted on the dynamometer scale beam and the engine speed.

~ I rd I $-, 1) I CH c cg;: | |Fuel Tank d Cd o ~ Extension C z g Extension KMercoid n Fuel Bottle r I o rStandard \ ~ Weights (12 r) I 4- IControl Box Revolution Time Counter | i Counter II Fuel Supply to Engine o 0 Figure 11. Automatic Fuel Weighing and Revolution Counting Devices Figure 11. Automatic Fuel Weighing and Revolution Counting Devices

-26Gas Pressure Indicator: A cathode ray indicator was used to obtain a pressure-time diagram of the gas pressure in the cylinder. The electrical circuit consisted of: 1. The pressure pick-up unit. 2. The degree marking unit. 1. The Pressure Pick-Up Unit: This was of the two catenary shaped diaphragm type. Its function was to convert the cylinder pressure variations into corresponding voltage variations which were applied to the input of a bridge-amplifier. The output of the amplifier was fed, through a switch, to channel A of a dual beam oscilloscope, figure (13). The trace obtained on the screen was photographed by a polaroid camera attached to the oscilloscope. 2. The Degree Marking Unit: The degree marks were produced by a steel disc, 20 inches diameter, 1/8 inch thick, mounted on the engine flywheel. The rim of the disc was slotted at 3~ intervals, with deeper slots at 45~ intervals, figure (12). A magnetic pick-up was mounted on the flywheel casing, with its pole close to the rim. The alternating voltage generated by the rotation of the disc was applied to channel B of the dual beam oscilloscope of item 1. The corresponding figure obtained on the screen consisted of a serrated line across the horizontal diameter of the screen. Every three and forty five degrees were thereby marked, and one of the deep 45~ slots in the disc was aligned at the top dead-center, then the crank angle positions along the indicator diagram were directly determined.

Magnetic Pick-Up Unit Amplifier Unit Steel Di Mounted on_______________ \ Engine Fly Wheel. Channel B -4 Slot a _ Dual-Beam Cathode-Ray Tube Unit Figure 12. The Degree Marking Unit.

Bridge-Amplifier Combustion _* Chamber Temp- _ * erature Signal " L _ l | | l IChannel A -F —- I., _, — Dual-Beam Switch < Cathode-Ray Tube Unit ru 45 v. Bri e-Am ii r Gas Pressure - SignalFigure 13. Layout of Electrical Circuits for Pressure and Temperature Recording.

-29Combustion-Chamber Thermocouple: To measure the rapid temperature changes of the combustionchamber-inside-surface a special Nickel-Steel thermocouple, manufactured by the Detroit Controls Corporation, was used. As shown in figure (14), it consisted of a: "Tyni-couple" basic unit, "Tyni-couple" mount, receptacle, plug, and coaxial cable. As described by the manufacturer, the basic unit, figure (15), was constructed by taking a 0.010" diameter nickel wire, applying a 0.00015" dielectric surface coating by oxidation, and swaging the wire into a slightly tapered hole in a small steel supporting body, thereby giving the assembled unit the ability to withstand very high pressures. The tip of the body and end of the wire were polished and plated with a 0.00025" thickness of nickel. The interface between the nickel plate and the steel body thus became the thermocouple junction. The basic unit was held in a steel mount, figure (14). The mount was equipped with external threads which were used to mount the unit on the cylinder head as shown in figure (16) and (17). Since the only electrical contact desired between the nickel and the steel was that existing at the hot junction, the internal nickel lead wire was insulated throughout its length from the steel parts of the thermocouple by a vinyl tubing. The nickel wire was soldered to the fitting in the receptacle. Shielded copper wires were used to carry the signal to the bridge-amplifier, then to the switch and the cathode ray oscilloscope. The layout of the temperature recording system being employed in the tests is shown in figure (13). The oscilloscope trace was photographed with a polaroid camera. The reference junction of the thermocouple circuit was at the ambient temperature.

Tyni-couple Mount Receptacle Plug Tyni-couple -Mounting Threads Coaxial Cable Figure 14. "Tyni-Couple" Assembly (Scale: Double Size)

-51Vinyl Tubing "Tyni-Couple" o061 Swaged Into Body 32.35" A ^ 1-.125" h a.0625" Oxidized Nickel Wire Swaged Into Body Figure 15. The Basic "Tyni-Couple" Unit

Injection Nozzle Energy Cell on Nozzle / -- Combustion Chamber 1" / / J h "Tyni-Couple. Cooling Water Jacket Cylinder Head "Tyni-Couple" Mount Figure 16o Sectional Plan of Cylinder Head Showing the Combustion-Chamber Thermocouple.

OF C. CHAMBER THERMOCOUPLE ig / CombustionC Chamber Figure 17- Sectional Elevation of Cylinder-Head Showing Position of Combustion-Chamber Thermocouple

i I i' i I' i' I..... I I i I i CALIBRATION CURVE FOR GUN STEEL - NICKEL 1200 - THERMOCOUPLE - - (REFERENCE TEMP - 200 C) z AMBIENT 5 1000 aE /DETROIT CONTROLS CORP. RESEARCH DIVISION w REDWOOD CITY, CALIFORNIA 0 n 800 W 600 a I / I U. 400 o: I - / Q 200 I0 4 8 12 16 20 24 28 32 VOLTAGE (.V.) FIG.18 CALIBRATION CURVE FOR THE COMBUSTION CHAMBER THERMOCOUPLE (Copy from figure delivered by monufacturer)

-35Piston at Bottom Dead Center Figure 19o Thermocouples Positions on the Cylinder-Liner Walls y ^ S S-tes===== ^ ===-______I~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~ -^g^:-g |^^^^| ~~Y~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~~

-36For the thermocouple, the theoretical response time, (time to reach 62.3% of imposed surface step change) = 4.5 microseconds. This time equals that required for the crank-shaft to rotate 0.034 crank-angle at an engine speed of 1200 R.P.M. Figure (18) shows a copy of the calibration curve for the thermocouple, as furnished by the manufacturer. Engine Walls Thermocouples: The thermocouples measured the surface temperatures of the liner, exhaust manifold, and the engine outer surface. The liner wall coolant side temperature, was measured at three points in a vertical plane perpendicular to the crank shaft. One of the thermocouples measured the wall temperature near the T.D.C.; the other at the B.D.C.; and the third, midway between the other two. The positions of these thermocouples are shown in figure (19). The exhaust manifold wall temperature was measured on the coolant side at a point midway between the exhaust valve and the manifold exit from the cylinder head. This temperature and the liner temperatures were used to calculate the coefficients of heat transfer from the gas to the walls, and from the walls to the cooling water. The engine outer surface temperature was measured at five points: two on the cylinder head, one on the cylinder barrel, and two on the crank case. They were connected in parallel to give the average surface temperature.

IV. EXPERIMENTAL PROCEDURE Procedure of Tests: The engine was started, and after the necessary warming up period, the following data were taken for each run: Air pressure in the surge tank of the intake system P Air pressure at the inlet tap of the orifice meter Pa Differential pressure across the air orifice LPa Differential pressure across the cooling water orifice APc Air temperature at inlet to orifice meter Ta Air temperature at the intake manifold T Lubricating oil temperature in crank case T Wall temperature of the exhaust manifold Twex Wall temperature of liner at T.D.C. T T1 Wall temperature of liner between T.D.C. and B.D.C. T T2 Wall temperature of liner at B.D.C. T Engine outer surface temperature Ts Total rise in cooling water temperature ATc Cooling water temperature at inlet to the engine Tc Cooling water temperature at inlet to the cylinder head Tc Cooling water temperature at exit from the engine T c3 Exhaust gas temperature Texh. Time for consumption of 1 lb. of fuel tF Revolutions of the crank shaft for the same period of time, tF NtF Brake load WD -57

-58Pictures of the following traces: Gas pressure and crank angles for the whole cycle. Gas pressure and crank angles for the compression and expansion strokes. Calibration of the gas pressure trace. Combustion chamber wall temperature and crank angles for the whole cycle. Combustion chamber wall temperature and crank angles for the combustion period. Calibration of the temperature trace. A sample of these data is given in Appendix A, and a sample of the pictures is shown in figures (20), (21), (22) and (23). A few runs were made with the back pressure tank between the engine and the main exhaust conduit. These were to find the effect of the back pressure on the brake mean effective pressure. Other data taken but not used in the calculations or results included lubricating oil pressure, lubricating oil temperatures before and after the cooler, current in the electric circuit of the air heater, air pressure after the reducing valve, and the blow-by rate of flow. Many of these were recorded for precautionary measures since by supercharging, the engine was overloaded beyond the design conditions. The usual friction run was made at the end of each day's testing. For the brake horse power the formula was: W N B.H.P. = B 4000 where N = revolutions per minute 4000 = constant for the dynamometer

-39The indicated horsepower was obtained by adding, the brake horsepower at any given speed, to the friction horsepower at the same speed. The cooling water temperatures at the inlet and exit from the engine were measured by two thermocouples connected to an ordinary "Leeds and Northrup" potentiometer. The difference between these two temperatures was also measured by two other thermocouples, connected to a potentiometer of the same type, but with greater precision reading temperatures down to 0.03~F. Four readings were taken for each run, and their average was used to calculate the heat losses to the cooling water. The water temperature was also measured at three of the four openings between the cylinder and cylinder head. During the first few runs, it was found that these three temperatures were almost equal. It was decided then, to connect the thermocouples in parallel, to get an average reading. The results given in Appendix A show the total rise in the cooling water temperature, the temperatures at the inlet to cylinder head and at the exit from the engine. The Tests Covered the Following Ranges: F ratios from.0134 to.055 A Air manifold pressures up to 45 inches mercury. Air manifold temperatures up to 204~F. Engine speeds from 560 to 1770 R.P.M. Since the engine was designed to run under natural aspiration conditions, and to avoid troubles due to overloading by supercharging, the following limits were considered: Exhaust temperatures not to exceed 1000~F. Gas peak pressure in the cylinder not to exceed 1200 lbs. sq. in. The engine at the end of the tests was in a satisfactory condition, without any sign of failure in any of its parts.

-40Calibration of the Combustion Chamber Thermocouple: The calibration curve for the nickel-steel thermocouple, which was used to measure the combustion chamber wall temperature, was furnished by the manufacturer [ Detroit Controls Corporation ], and is shown in figure (18). To calibrate the temperature trace on the oscilloscope screen, the engine was stopped, and while the cylinder was still hot, the voltage produced by the thermocouple was measured by a precision potentiometer. At the same time a picture of the calibration trace was taken. The trace as shown in figure (23), is in the form of two parallel dashed lines; the vertical distance between them being equivalent to the voltage produced by the thermocouple. The reading of the potentiometer was in millivolts, and was converted to degrees fahrenheit by using the calibration curve fugure (18). The scale of the temperature trace was then calculated from the calibration trace and the reading of the potentiometer.

Trh it.;~ *r V: ~............... ~ a ~~ ~ Y 2,,." c~ e:~ ~

;0.,,. -, ~3 n v ci;:

.... 4..... 5 i X. ~a,~ lj I Ierrap@v/.. |Qs i;be F*;gl ~:~ {3 f Ct A. Wall. tem)peratu.re over the whole cycle and the atmospheric tLemperaturo.de (With 5 mpl ifte ) B. Wall temperature over the whole cycle (With 2 amplifiers) C, Wall temfjlrature e over the whole cycle and the atmlnosheric t femperf atxre. (With 2 amplit:.fl ers) Fi.gura.*e 22. Comlbuastion Chambex Sur f.ace T.rans.ientr en....Tex:'rature.

-44lb s Pressure,(right): Deflection corresponds to 231.3 sq. in. Temperature,(left): Deflection corresponds to 33.1~F a. Deflection with 2 amplifiers (the two middle lines). b. Deflection with 3 amplifiers (the two outside lines). Figure 23. Pressure and Temperature Calibration Traces.

V. EXPERIMENTAL RESULTS. The results shown in this section covered the following range: Series A. Runs At Variable Manifold Pressures: Figures (24-32): Average Speed = 800 R.P.M. Average Manifold Temperature = 80~F Manifold Pressures = 30" Hg - 45" Hg. F Ratios:.0134 -.055 A Series B. Runs At Variable Manifold Temperatures: Figures (33-38): Average Speed = 1200 R.P.M. Manifold Temperatures = 80 - 204~F Average Manifold Pressure = 36" Hg. E Ratios:.0213 -.0468 A These results are also given in tables (3-11). Other runs were made at different engine speeds and were used only to check the temperatures of the combustion chamber calculated from equation [2.14]. The results of these runs are shown in tables (12 and 13). Other set of runs was made before supercharging the engine to check the engine conditions while it was naturally aspirated. These are not of interest in the present investigation and their data are not given in this report. -45

-46SERIES A EFFECT OF F RATIO ON B.M.E.P. FOR VARIOUS MANIFOLD PRESSURES A Average Speed: 800 R.P.M. Average Air Temperature: 80~F O Manifold Pressure: 30" Hg & Manifold Pressure: 35" Hg 150 + Manifold Pressure: 36" Hg x Manifold Pressure: 9" Hg * Manifold Pressure 42" Hg 140_ 0 Manifold Pressure 45" Hg 120 110 O| |~ ~ LINES OF CONSTANT 3 100 FUEL CONSUMPTION E- 7 / / ^/ LBS.FUEL | 80.5 HR. 60 50 _ 2.5 40 2.0 30 20 I0 0 0.01 0.02 0.03 0.04 0.05 0.06 FRATIO A FIGURE 24

-47SERIES A EFFECT OF RATIO ON I.M.E.P FOR VARIOUS MANIFOLD PRESSURES AVERAGE SPEED:800 R.PM. AVERAGE AIR TEMPERATURE: 80~ F. z 170 o MANIFOLD PRESSURE: 30" Hg d ^ A 33 1: C60 +: 36 / W. a 16x I:39 w 0 gI 1I 42.sj, II: 42 / / / 150":45,W, 140 z t 1J / x 130 w I — 120 _ 10 _ F RATIO 10 90 80 + 70 60 50 40 30 20 10 0 0.01 0.02 0.03 0.04 0. 05 0.06 F RATIO FIG. 25

-48Series A _ I I, I I.M.E.P. ~ 14.7 Tm VS. F RATIO L ^Pm 50JP A FOR VARIOUS MANIFOLD PRESSURES Average Speed: 800 R.P.M. 160 Average Air Temperature: 80~F 0 Manifold Pressure: 30" Hg g~150Q A Manifold Pressure 335" Hg + Manifold Pressure: 36' Hg ~~~140 ~X Manifold Pressure 39" Hg * Manifold Pressure 42t" Hg 1 Manifold Pressure: 45" Hg 120 _ o / ~ 0 -H 110 #6, 100 / + E-4 90 x - 80, +00 70 P4 o: 60 _ 50_ 40-r 30 I0_ I I I I I 0 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO A FIG.26

-49Average Speed: 800 R.P.M. Average Intake Temperature: 80~F 0 Manifold Pressure: 30" Hg A Manifold Pressure 33" Hg + Manifold Pressure 6 36" Hg x Manifold Pressure:39" Hg * Manifold Pressure 42" Hg o Manifold Pressure:45" Hg 0.9 0.8 A 0.7 0.6 0.4 SERIES A B.S.F.C. VS. F RATIO FOR VARIOUS MANIFOLD PRESSURES 0 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO FIG. 27

-50Average Speed 800 R.P.M. Average Air Temperature 80~F 0 Manifold Pressure 30" Hg A Manifold Pressure 35" Hg 90 + Manifold Pressure:6" Hg w Manifold Pressure 39' Hg * Manifold Pressure: 42" Hg * Manifold Pressure 45" Hg 80 70 60t 50 o | / / / SERIES A / / ~/ EMECHANICAL EFFICIENCY VS. F RATIO 40 FOR VARIOUS MANIFOLD PRESSURES 30 20 I I I I L 0 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO A FIG. 28

-51SERIES A HEAT LOSSES TO COOLING WATER VS. F RATIO A FOR VARIOUS MANIFOLD PRESSURES Average Speed: 800 R.P.M. Average Air Temperature: 80~F O Manifold Pressure 30" Hg 30000 a Manifold Pressure 35" Hg 4 Manifold Pressure 36" Hg X Manifold Pressure 39" Hg Manifold Pressure 42" Hg 26000 x + D 22000 1000 I_ /I II 0 0.01 0.02 0.03 0.04 0.05 0.06 14000 co 0 OF 10000 6000 2000 0 0.01 0.02 0.0:3 0.04 0.05 0.06 F RATIO A FIGURE 29

SERIES A HEAT LOSSES TO COOLING WATER VS. INDICATED M.E.P. FOR VARIOUS MANIFOLD PRESSURES 30000 Average Speed: 00 R.P.M. Average Air Temperature 80~F 0 Manifold Pressure 30" Hg X ~ Manifold Pressure 55 Hg. 25000 + Manifold Pressure: 36" Hg + + Pim~ |X Manifold Pressure 39" Hg E-^ * I~ Manifold Pressure 42" Hg G Manifold Pressure 45" Hg L 4 ~ 20000 X A /X 0 +J Co E0i 0 + 15000 0 oooo fo E40 10000 0 50001 0 20 40 60 80 100 120 140 160 INDICATED M.E.P. lbs./sq.in. FIGURE 30

-53SERIES A EXHAUST GAS TEMPERATURE VS. F RATIO FOR VARIOUS MANIFOLD PRESSURES A Average Speed 800 R.P.M. Average Air Temperature 80~F O Manifold Pressure 30" Hg D Manifold Pressure 5335 Hg + Manifold Pressure 36" Hg X Manifold Pressure 39" Hg * Manifold Pressure 42" Hg o Manifold Pressure 45" Hg 1000 oo / X I 600 EJ AA 400 - 200 0. 02 0.03 0I04 0I05 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO A FIGURE 31

Average Speed 800 R.P.M. Average Air Temperature 80~F O Manifold Pressure 30" Hg A Manifold PresLure 33" Hg + Manifold Pressure 36" Hg X Manifold Pressure 39" Hg 42 ~ Manifold Pressure 42" Hg 0 Manifold Pressure 45" Hg X 40 H 380 36 34 k~f FOR VARIOUS MANIFOLD PRESSURES 2;8 0 20 40 60 80 I 00 120 140 160 INDICATED MEAN EFFECTIVE PRESSURE lbs./sq. in. FIGURE 32

-55SERIES B BRAKE MEAN EFFECTIVE PRESSURE VS. F RATIO FOR VARIOUS INTAKE AIR TEMPERATURES Average Speed 1200 R.P.M. Average Air Pressure: 36" Hg O Intake Air Temperature: 80~F X Intake Air Temperature: 140~F + Intake Air Temperature: 200~F 140 120 100 80 40 20 0 I I I I I I 0 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO A FIGURE 55

-56SERIES B 180 - INDICATED MEAN EFFECTIVE PRESSURE VS. F RATIO A FOR VARIOUS INTAKE AIR TEMPERATURES 160- Average Speed: 1200 R.P.M. Average Air Pressure 36" Hg O Intake Air Temperature: 80~F X Intake Air Temperature: 140oF + Intake Air Temperature: 200'F 140 K/ /// 120 X 1 00 80 60 40 20 0 0.01 0.0 2 0.02 0 0.04 0.05 0.06 F RATIO A FIGURE 34

-57SERIES B.M.E.P. Tm VS. F RATIO FOR VARIOUS INTAKE AIR TEMPERATURES ~L 5U A Average Speed 1200 R.P.M. Average Air Pressure: 36" Hg O Intake Air Temperature: 80~F X Intake Air Temperature: 140~F 160 + Intake Air Temperature: 200~F 140 g~~~~~ ~/ 0 0 t, 10oo ~ I H 80 / * yX 60 40 20 0 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO A FIGURE 55

-58SERIES B BRAKE SPECIFIC FUEL CONSUMPTION VS F RATIO FOR A VARIOUS INTAKE AIR TEMPERATURES Average Speed 1200 R.P.M. Average Air Pressure: 36" Hg 0 Intake Air Temperature: 80'F X Intake Air Temperature: 140~F + Intake Air Temperature 200~F 0.7 0.6 00 0.4 0.3 0 0.01 0.02 0.03 0.04 0.05 0.06 F.RATIO A FIGURE 56

30000 25000 | 20000 0 0 150000 SERIES B vu Pq HEAT LOSSES TO COOLING WATER VS. INDICATED M.E.P. CO 0FOR VARIOUS INTAKE AIR TEMPERATURES Average Speed 1200 R.P.M. Average Air Pressure 56" Hg O Intake Air Temperature: 80~F 5000 X Intake Air Temperature: 140~F + Intake Air Temperature: 200~F o ~0~ —- 20 40 60 80 100 120 140 160 INDICATED M.E.P. lbs./sq.in. FIGURE 37

-60SERIES B EXHAUST GAS TEMPERATURE VS. F RATIO 1000- K FOR VARIOUS INTAKE AIR TEMPERATURES 0 Average Speed: 1200 R.P.M. / X 900 Average Air Pressure: 36" Hg 0 Intake Air Temperature: 80F / X Intake Air Temperature: 140~F X + Intake Air Temperature: 200~F O 800 + | 700 - EQ / 500 400 300 200 L lI ---- 0 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO A FIGURE 58

VI. DISCUSSION OF EXPERIMENTAL RESULTS Series A. Runs At Variable Manifold Pressures: Effect of F ratio on B.M.E.P. for Various Manifold Pressures Figure 24: The B.M.E.P. of the engine increases with increaseing F ratios and supercharging pressures. The increase in the power output is mainly due to the increase in the amount of fuel used by the engine. This can be achieved either by increasing the fuel-air ratio at a constant intake manifold pressure, or by increasing the intake manifold pressure at constant fuel-air ratio. However, the power output obtained by using equal amounts of fuel at high manifold pressures is more than that obtained if the same amount were to be used without supercharging; therefore it would be better to increase the output by pressure charging rather than by increasing the F ratio. This fact A is illustrated by cross plotting the lines of constant fuel consumption per hour on the same figure. These lines show also that supercharging results in a rise in the B.M.E.P. up to a limiting pressure of about 40" Hg, beyond which there is no further increase; eventually there is a decrease. Some factors which might affect this limiting pressure are the fuel injection system design and timing, type of combustion chamber, and the inlet and exhaust valve timing. Effect of ~ ratio on I.M.E.P. at Various Manifold Pressures, Figure 25: The engine I.M.E.P. curves follow the same shape as those for the B.M.E.P. But a sharp rise in the I.M.E.P. is noticed by comparing the atmospheric aspirated output with that obtained at -.61

-6233" Hg manifold pressure. This is mainly due to the better scavenging efficiency resulting from any slight supercharging. The sharp rise in power output is not noticed in the B.M.E.P. curves, due to the relatively poor mechanical efficiency of the engine at low supercharging pressures. [ I.M.E.P. x 14.7 x I ] vs F ratio for Various Manifold Pressures, m 540 Figur2 26: It is noticed that practically all the points of figure (25) are well correlated by a single curve except for some of the runs at naturally aspirated conditions. This departure from the common curve is due to the poor scavenging efficiency of the engine at these conditions. From this figure it can be concluded that for any fuel-air ratio the power output varies in direct proportion to the weight of the air used by the engine. F B.S.F.C. vs F for Various Manifold Pressures, Figure 27: The better brake specific fuel consumption of the supercharged engine is due to its higher mechanical and indicated thermal efficiencies. Also, as will be shown later, the heat rejected to the cooling water is less, thus more of the fuel heat is left to be converted into net engine output. Mechanical Efficiency vs F ratio for Various Manifold Pressures, Figure 28: The mechanical efficiency of the supercharged engine is higher than that of the naturally aspirated engine simply because the mechanical losses increase very little with increased engine output and percentage-wise become much smaller.

-65F Heat Losses to Cooling Water vs - ratio for Various Manifold Pressures, Figure 29: The heat losses to the cooling water increase with both fuel-air ratio and intake manifold pressures. These losses consist of the direct heat loss from the gases, and the friction heat. The direct heat loss is a function of the gas temperature and the coefficient of heat transfer, which is discussed in detail in Chapter VII. Heat Losses to Cooling Water vs I.M.E.P. for Various Manifold Pressures, Figure 30: From this figure it can be concluded that the heat losses at any supercharging pressure and - ratio is a function of the A I.M.E.P., all other conditions being held constant. Also for the same B.M.E.P. the heat losses decrease with supercharging. F Exhaust Gas Temperatures vs F ratio for Various Manifold Pressures, Figure 31: The exhaust gas temperature is mainly a function of the fuel-air ratio, but for manifold pressures higher than 33 inches mercury the exhaust gas temperatures are slightly higher than those for normal atmospheric operation. This is due to the relatively higher temperatures reached in the cycle with boosted inlet pressures. Indicated Thermal Efficiency vs I.M.E.P. for Various Manifold Pressures, Figure 32: Although the points for each manifold pressure do not fall on a smooth curve, yet from the average curves drawn for the different pressures it can be noticed that at high I.M.E.P.'s the thermal efficiency of the supercharged engine is better than that of the naturally aspirated engine. This is due to the fact that the supercharged engine operates at a lower fuel-air ratio and with a better scavenging efficiency. The slight drop in the indicated thermal efficiency of

-64naturally aspirated engine, as the I.M.E.P. increases, is caused by the high fuel-air ratios used and the high gas temperatures reached at the beginning of the compression stroke. Series B. Runs At Constant Manifold Pressures And Variable Manifold Temperatures: Effect of F ratio on Power Output for Various Manifold Temperatures, Figures 33 and 34: The heating of the air before it enters the engine causes a drop in the power output. This is mainly due to the drop in the mass-rate of air flow caused by an increase in the specific volume of the air at the time the inlet valve closes. As the cycle starts from a higher initial temperature the gas temperatures throughout the cycle are increased, causing some decrease in the indicated thermal efficiency, accompanied by an increase in the heat lost to the cooling water. [I.M.E.P. x TM ] vs F ratio at Various Manifold Temperatures, Figure 35: This figure confirms the conclusion reached from figure (26), F i.e., with constant- ratio, the power output varies in direct proportion to the wieght of air used by the engine. B.S.F.C. vs- for Various Manifold Temperatures, Figure 36: A The poor fuel economy at higher manifold temperatures is caused by the drop in the thermal efficiency as explained in the discussion for figures (33) and (34). Heat Losses to Cooling Water vs F I.M.E.P. ratio for Various Manifold Temperatures, Figure 37. The heat losses to the cooling water increase with an increase in the manifold temperature. This is due to the high temperatures

-65and coefficients of heat transfer reached during the cycle. F Exhaust Gas Temperatures vs - ratio for Various Manifold Temperatures, Figure 38: The exhaust gas temperature is a function of the fuel-air ratio for all intake manifold temperatures.

VII. HEAT-TRANSFER ANALYSIS OF THE EXPERIMENTAL RESULTS. The purpose of this analysis is to apply equation [2.14] and [2.17] to calculate the combustion-chamber-wall inside-surface temperature and the thermal loading on the cylinder walls, and to check the values calculated, with those measured. This analysis was made in the following steps: 1. Study of the effect of the manifold pressure and temperature, and the mean piston speed, on the gas mean effective temperature, and coefficient of heat transfer. 2. Evaluation of some constants for the engine. 5. Calculation and check on the wall temperature. 4. Calculation and check on the thermal loading. 1. Effect of Pm, Tm and S on TM.E. and M: TM.E. and afM are the main factors affecting the process of heat transfer from the gas to the walls. The following procedure was followed to find an equation for them in terms of Pm, Tm, S and the I.M.E.P. a. Calculation of the gas pressure all around the cycle from the pictures taken for each run. A sample is shown in figure (20). b. Calculation of the gas temperature all around the cycle from the weights of air and fuel consumed for each run. c. Substituting in equation [2.1], the values of the pressure and temperature previously caluclated, and -66

-67the value of the mean piston speed measured during the tests; to get the coefficients of heat transfer. d. Plotting the values of ag against the crank angles, (as in figure 52) and by graphical integration getting the mean value of ag, known as aM. e. Calculating CgTg all around the cycle by multiplying the instantaneous value of ag with the corresponding value of Tg. f. Plotting CgTg against the crank angles, (as in figure 53) and by graphical integration getting the mean value of agTg, known as (cXgTg)M. g. Using equation [2.8] to calculate the gas mean effective temperature over the cycle. A sample of these calculations is given in Appendix A. This cycleanalysis was made for many runs covering the following range: Pm: 28.75 - 44.25 inches Hg. Tm: 73.5 - 204 ~F 9: 11.3 - 17.97 ft. sec. N: 775 - 1231 R.P.M. The results of these analyses are given in tables (16) and (17). The best correlation for these results was found by plotting: cqM 1 CM x 1 vs I.M.E.P. Figure (39) and EaM TM.E. vs I.M.E.P. Figure (40) EJS Tm

a. 1 vs. I.M.E.P. 03.8NS 3AfjPmTm ~~~~~~~0.78~~0 0.7 o of 0 0 00 0.5 0.4 I I I I I I I,I 0 20 40 60 80 100 120 140 160 I.M.E.P. lbs./sq.in. FIGURE 39

50- aM. TM.E. VS. I.M.E.P. aJ Tm o 40 ~- O 10 -- I I I I I I I I O0 20 40 60 80 100 120 140 160 FIGURE 40

TM.E. calculated from cycle. analysis R aM calculated from cycle-anolynsi B.T.U. ).^ 0 0) 5O f ^ Z hr.sq.ft. 0F 0 0 0 0 0 0 0 0 0 0 N to 0 0 0 0 0 0 0 0 0 0 0 0 0 0 0 00 0 -- 0 o -- K O N 0 0 - 0 - 0 0 a-,o m o - 00 0e I1~ - 90.4 0 o N 90 ) 0 - I Il I I I,I I I;D a

-71The relation between these variables were found to be of the form: M x 1 = 0.48 + 0.00157 (I.M.E.P.) 3 3 Pmm and aM x TM.E. = 14.71 + 0.233 (I.M.E.P.) 3-S Tm From these two equations: TmTME 14.71 + 0.233 (I.M.E.P.) TM'E = ~ P —T 0.48 + 0.00157 (I.M.E.P.) [7.1] and aM = 3,S x 3P [0.48 + 0.00157 (I.M.E.P.)] [7.2] The values of TM.E. and XM calculated from equations [7.1] and [7.2] were plotted against those found from the cycle-analysis, in figure (41). The maximum errors in TM.E. and aM are 7.2% and 4.5% respectively. 2. Evaluation of Some Constants for the Engine: a. Constant C in equation [2.11]. ac c..... [7-31 w 0.6 c0.4 (- ) (-jk) Wc was taken from the cooling water flow rates measured for each run. c,>, and k were evaluated at the average temperature of the cooling water. ac was calculated from: Q, the heat carried by the cooling water; AW. C the area of cylinder

-72walls in contact with coolant; and the mean temperature difference between the walls and the water. The area of cylinder walls, AW.C. consists of: (i) the barrel area (ii) the combustion chamber area (iii) the exhaust manifold area enclosed in the cylinder head. The mean average temperature for these areas was calculated by adding the multiple of each part by its temperature, then averaging the sum over the whole area. Since the flow of water past the surface was a mixture of parallel, counter, and cross flows; ac was calculated in terms of the arithmatic mean temperature difference between the surface and water. A sample of these calculations is shown in Appendix B. The results of similar calculations for other runs are tabulated in tables (14) and (15), and were plotted in figure (57). From figure (57) C for the engine = 5.15 [7.4] Thus ac =.15 (Wc)0.6 (c)0.4 b. Area factors for the combustion chamber: The shape of the combustion chamber is shown in figures (16) and (17). To evaluate the area factors,

-73the combustion chamber wall was considered as cylindrical, with an inside diameter of 2 inches, and an outside diameter of 2.56 inches. Mean area factor am: ~m 2.28 am Ag = E -.877 [7~51 Cooling area factor ac: A.- 2 ac.8=.28_ a. A 2. 6 -.782 [7.61 Wall equivalent thickness x: D2 x = rm lnD1 2 + 2.56 2.56 = 256 x in 26 0.282 inch. 2 2 [7.7] c. Thermal conductivity of the metal: k = 27 B.T.U. [7.8] vw ~ hr. sq. ft. (~F) ft 3. Calculation and Check on the C. Chamber Wall Temperatures: By equating the values of: TM.E., equation [7.1]; OC, equation [7.2] and the constants, equations [7.4] to [7.8]; in equation [2.14]; the values of Tw.g. could be calculated from: PmTmS, I.M.E.P., Wc, Tc3 and the engine constants. Using this procedure, Tw.g. was calculated for several runs covering the following range of operating conditions: * (Reference No. 53)

-74Pm: 28.75 - 44.7 inches Hg. T: 72 - 204 ~F m N: 789 - 1769 R.P.M. lbs. I.M.E.P.: 51.3 - 152 lbs. sq. in. The results are shown in tables (18) and (19). These calculated temperatures were compared with those measured by the thermocouple and were found to be in fair agreement. The maximum error is 12.7%. The values measured are shown also in tables (18) and (19). The deviation of the measured temperature from that calculated is due to the following experimental errors: a. The time when the picture of the wall-temperature was taken was not exactly the same time the cooling water was measured, during which a change in the water temperature might have taken place. The change in the water temperature was within + 5~F during the run and a result of the thermostatic control being not very sensitive. b. The indicated mean effective pressure used to calculate the wall temperature was evaluated by adding the friction load measured at the end of the day's tests to the brake load for the run. The friction load, being a function of the lubricating oil temperature, was not constant for all the runs of the day, and caused an error in the indicated mean effective pressure. c. The cooling water temperature Tc3, used in the calculations, being lower than the water temperature near the combustion chamber wall, caused the majority of the calculated wall temperatures to be lower than those measured by the thermocouple.

-754. Calculation and Check on the Thermal Loading: By equating the values of: TMHE. equation [7.1]; o, equation [7.21 and the constants, equatio n[7.4] to [7.8]; in equations [2.16], [2.17] and[2.18]; the thermal loading could be calculated from PmP TMP So I.M.B.P., W (cl + To) and the engine constants. Equation [2.18] was simplified by assuming the mean effective area equal to the mean arithmatic area, therefore AM - A.ch.+ Ap.T. + 2 Aint man,+ Aexh man. +[ 4 ] - 2.5 + 15.9 + 37.1 + [ 26.6 100.7sq.ins The results of the calculations are shown in Tables (20) and (21) as compared with the measured cooling water heat losses. The calculated thermal loadings are in general less than the measured heat losses which included part of the friction heat. The deviation between the calculated and measured values increases at high engine speeds due to the increase in the friction heat. Another factor which affects the measured losses is the heat losses from the outer surface of the engine.

-76VIII. INVESTIGATION ON THE EFFECT OF AFTERCOOLING To investigate the effect of aftercooling on the operation of the turbocharged engine, the following calculations were made, based on the conclusions reached from the experimental work. The power output, the maximum temperature of the piston crown and the intensity of thermal loading were calculated for different degrees of aftercooling. The arrangement on which calculations were based is shown diagrammatically in figure (42). The conditions of operation assumed were as follows: Compressor Efficiency = 80% Atmospheric Conditions Pressure = 30 inches Hg. Temperature = 80 ~F Drop in air pressure between the compressor outlet and the engine manifold = 0.5 lb sq. in. Aftercooler Effectiveness = 0, 50% and 100% Method of Calculation: a. Calculation of I.M.E.P. for different aftercooler effectiveness: After being compressed in the turbocharger, the air will be at a temperature higher than that of the atmospheric. If an aftercooler is used between the compressor and the engine, the air temperature will drop to a value depending on the aftercooler effectiveness; and the power output will be greatly affected. It was found from the results of the experiments that the I.M.E.P. is inversely proportional to the absolute manifold temperature. Thus the power of the turbocharged engine, at

-77Intake Exhaust Manifold r Manifold 7 -_ ThrTurbine Exhaust & Turbine ~ T 1 - -Air Intake T1, P1 rn Engine T1^ Pi _ _'\ 2 ) Compressor Aftercooler Figure 42. Diagram of Assumed Turbocharged C.I. Engine

-78any manifold pressure and with any aftercooler effectiveness, can be calculated by multiplying the I.M.E.P. from the test-results [figure (25)] by a temperature factor 5 The test-results were Tm considered (after being corrected to a manifold temperature of 540~R and manifold pressures of 30 to 45" Hg) to represent the engine performance with 100% aftercooling effectiveness. These are shown in figure (43). b. Calculation of the maximum temperature of the piston-crown: For these calculations, the following constants were evaluated: x: the thickness of the piston-crown = t inch kw: coefficient of thermal conductivity = 27 B.T.U. hr. sq. ft. (~F ft Ta: the temperature of the hot gases in the crank case = 160~F aa: coefficient of heat transfer from the bottom of the crown to the hot gases, calculated from equation [2.1] with P = 14.7 T = 460 + 160 = 620 R ft. S: at 800 R.P.M. = 11.672 se. sec e -a.a = 12.22 B.T.U. hr. sq. ft. ~F c. Calculation of Intensity of thermal loading on the combustion chamber walls: This was calculated from equation [2.15]: q = U (TM.E.-TC) U was calculated from equation [2.16] by using equations [7.2], [2.11] and [7.4] to [7.8] and TM E was calculated M. equation [7.1. from equation [7.1].

-79A sample calculations for the effect of aftercooling on the engine performance, piston-crown maximum temperature and intensity of thermal loading is given in Appendix G. The results of similar calculations are given in tables (22), (23) and (24). They are plotted in the following figures: Figures (43), (60) and (61) for 100% aftercooler effectiveness. Figures (44), (62) and (63) for 50% aftercooler effectiveness. Figures (45), (64) and (65) for 0% (No aftercooling). Figures (43), (44) and (45) illustrate the engine performance for conditions of, constant manifold pressures, piston-top maximum temperatures, and intensities of thermal loads. The curves for the indicated mean effective pressures were calculated from figure (43) and the temperature factors. The other curves are cross plots from figures (60) to (65). At the end of this chapter calculations were made, to find the effect of aftercooling on the intensity of thermal loading, piston maximum temperature, and power output, at a manifold pressure of 45 inches Hg. The results are shown in figures (46), (47), (48), (49), and (50). Discussion of Results of Calculations: Performance with 100% aftercooler effectiveness, figure 43: This figure shows that increasing the power output at any supercharging pressure causes an increase in both the thermal loads and the piston temperatures. This is mainly due to the increase in the amount of fuel used per cycle. Also, the thermal loading and piston temperatures at high supercharging pressures and low A ratios, are lower than those at lower supercharging pressures and higher F A

-80Performance With 100% Aftercooler Effectiveness 3.170 AP 3800 34000 140, Co ~4Loads 120 T/U 120 hr.sq.ft. 1; I ~~~~~/ 9I/v y1 ^950 d2 026000 |.\//0C/ y 100 922000 75 90 40 F0 70 50 4-, 30 20 10 0 0.01 0.02 0.03 0.04 0.05 0.06 RATIO FIGURE 43

-81Performance With 50% Aftercooler Effectiveness L 380 A _ _160 Co 14 /0 3 ^ 130 _ Intensity of Thermal / 34000 / / / L Ioads - - I 120 B.oT.IU. / AF gw"~~ 12hr.sq.ft. 110 ^ ~030000 I 00^ / /^^ / Piston Temp. ^ 26000 /0( / / / zso / "F 90_:/7y o60~~~~ ~ 100800 soL -22000 / 7%0^ // 7/00 40_ 30 20 10 00 0.01 0.02 Q03 0.04 0.05 Q06 FRATIO FIG.44

-82Performance Without Aftercooling I ~Id~~I I It "P'160 K 150 -w cS 130 - Intensity of Thermal 0. I000 F Loads?/ r-.~~~ 120 BS^ BiTB.T.U. 120 Cd dOhr. sq.ft. Fr50 F 90 F 80 9/^ ^00 F ^'100 - 70 ~'"Piston Temp. _ o~ 8:^50 OF 90 ^ Jr,^ 80O F 750 F TO0~, 50 \%0 40 30 20 10 0 0.01 0.02 0.03 0.04 0.05 0.06 F RATIO Figure 45 A

-83ratios. Thus for equal power outputs it is better to operate the engine at the highest supercharging pressure, with the lowest F A ratio. This will also result in the best fuel economy. Performance with 50% aftercooler effectiveness, figure 44: This figure also shows that for any supercharging pressure, both the thermal loads and piston temperatures increase with the power output. In this case, the thermal loads are functions, only, of the I.M.E.P., for any F ratio or supercharging pressure. For the A same I.M.E.P. the piston maximum temperature drops with supercharging. For any brake mean effective pressure, it is still better to run the engine at the highest supercharging pressure, since the thermal loading and the piston temperature will be lower. Performance with 0% aftercooler effectiveness, figure 45: Here also both the thermal loads and piston temperatures increase with the engine output. Although the same I.M.E.P. could be reached at lower F A ratios with supercharging, yet the thermal loads increased and the piston maximum temperatures remained constant. However, for equal brake mean effective pressures, the piston temperatures will decrease with supercharging and the thermal loading will remain constant. For the turbocharged engines without aftercooling we have to choose either, the low efficiency and low mechanical loading of the low supercharged engine, or the higher efficiency but higher mechanical loading of the higher supercharged engine, without any

-84reduction in its thermal loading. Engine performance with different aftercooling effectiveness, figures 46 and 47: Both figures are for a constant manifold pressure of 45 inches mercury, with the variables arranged in two different ways. It can be noticed that: a. The wall temperatures and intensities of thermal loading are functions, not only of the fuel-air ratios, but also of the aftercooler effectiveness. For the same fuel-air ratio the thermal loading and wall temperatures decrease with better aftercooling. b. The fuel-air ratio required for equal power outputs decreases with better aftercooling, resulting in a reduction in the thermal loading and piston temperatures, and a better thermal efficiency. Effect of aftercooling on thermal loading: Figure (48a) shows the intensity of thermal loading with different supercharging pressures and aftercooling effectiveness, for an indicated mean effective pressure of 150 lbs. per sq. in., with a compressor efficiency of 80%. Curve for e = 100%: This is the case of supercharging and cooling to a constant manifold temperature of 540~R, (this is the condition of series "A" of the experiments). The thermal loading decreases continuously with supercharging. That is to say, the increasing manifold pressure has a reducing effect on the thermal loading, a pressure rise of 15" Hg causes a reduction of 7.7%. This reduction is not noticed in figure (30), because the manifold temperature and the friction heat were not constant for all the runs.

010 * 1100 -Effect of Aftercooler Effectiveness ~ ~ SC ~~~~~~~~~~~~~~~~~~~~~~~~~~~0.0,1 0 Pm ~ 45" Hg ^>< ^ \0000 o 1000 4,) 900 y00^ >-/ >^ -— >^3F00 8 0Ratio0 s " <^ / 6007 ^>C ^^ -— ><. ~36000 900 I Y3 34000 32000 QO y^^ ^ — 7^30000 PO 800. ^0^ 7^-^ ^ - 0 0 1 20 3 40 069 10 11026000 1 1 Q0, __^X X ^s~r o~nnn 1 Intensity of Thermal Loads;R B.T.U. 700 220 M hr.sq.ft. FIGUR 4-22000 0 * p 7 20000 )8000 600 O 500 I T I 0 10 20 30 40 5.0 60 70 80 90 100 110 120 130 140 150 160 170 Ibs. Indicated Mean. Efl-etive Pressure sq.in. FIGURE 46

-86| l~ l~ l i oI0\o \ Effect Of Aftercooler Effectiveness O o\o 0\0 Pm= 45" Hg 0 t 170 -/ (4& 15~0 -P ~ ~~ 0,,\ 1620 - co o 140 30 10 4-F 14Piston Max. Temp. 10 0 0 0 70 0 50QFI 7 70 260 10

-87Curve for e = 03: This curve represents the case of turbocharging without aftercooling. The increase in the thermal loading with turbocharging is mainly due to the increase in the manifold temperature, which has a much greater effect on increasing the thermal loading than the reducing effect of the increasing manifold pressure. An increase of 87~F due to turbocharging to 45" Hg, caused an increase of 22.3% in the thermal loading. If the thermal loading was not permitted to increase by turbocharging, it would be better to use low manifold pressures with better aftercooling than to use high manifold pressures and poor aftercooling. In other words, an effective cooler which might cause a bigger pressure drop, will be preferred to an uneffective cooler which will cause a small pressure drop. It is worthwhile to add here that, for an I.M.E.P. of 140, the thermal loading on the engine-cooling-system, is reduced by an amount twice as much as the heat removed from the air in the aftercooler. This is illustrated in figure (50). Effect of aftercooling on the maximum temperature of the piston-crown Figure (48,b) shows the piston-crown maximum temperature for different supercharging pressures and aftercooling effectiveness and an indicated mean effective pressure of 150 lbs. per sq. in., with a compressor efficiency of 80%. It indicates that the piston temperature drops with better aftercooling for any supercharging pressure.

-88Effect of aftercooling on power output: Figure (49) shows the increase in the power output by supercharging and different degrees of aftercooling. A horizontal dotted line representing an I.M.E.P. of 129 lbs. per sq. in. was drawn to intersect with the lines of e = 0, 50 and 100%. This line indicated that with better cooling, lower manifold pressures are required to give the same I.M.E.P. This means that, with an air pressure of 45" Hg at the exit from the compressor, it would be better, in the interest of lowering the thermal loading, to use an aftercooler with e = 50%, which might cause a pressure drop up to 3.8" Hg, rather than to exclude the cooler and keep the pressure at 45" Hg.

-89Effect of Aftercooling on Intensity of Thermal Loading and Piston Maximum Temperature I I I I I.~!; |I I ~t.compressor =80 % _~]^~~ I,, I.M.E.P. =150 bs./sq.in.' 44000 42000 40000 0 > 38000 %. 36000 34000 32000 FIGURE. 48. 1o10 e =o % Ca I I 0= OI)? looo 900 I I I I I 30 33 36 39 42 45 lbs. Manifold Pressure sq i FIGURE. 48b.

uw 150 140 130 -> 130 - C IL u.LL 120 0/ 0I' 110 100 I'3 AC)o I MANIFOLD PRESSURE inches Hg. FIG. 49

-91EFFECT OF AFTERCOOLING ON REDUCTION OF THERMAL LOADING AND PISTON MAX. TEMPERATURE Pm = 45" Hg COMPRESSOR EFFICIENCY =80 % 150 _ 100 0 50 0 5500 n 5000 - 4500 1000 1350000 2000 200 Heat Taken In Aftercooler 25 50 75 1003000 hr. 2000 _ P O 1500 1000 500 500 1000 1500 2000 2500 Heat Taken In Aftercooler I I I I B.T.U. 0 25 50 75 1oo % / hr. Aftercooler Effectiveness FIG. 50

IX. CONCLUSIONS AND RECOMMENDATIONS Conclusions: From the foregoing analysis it is concluded that: 1. The thermal loading and wall temperatures of a turbocharged compression ignition engine are primarily related to the indicated power output and are affected to a lesser degree by intake manifold conditions and the engine speed. Results of this investigation show that: a. Thermal loading and wall temperatures increase in direct propotion to the indicated mean effective pressure. b. For the same indicated power output: (i) Boosting the intake manifold pressure without heating the air will reduce the thermal loading and wall temperatures (ii) Raising the intake manifold temperature will increase the thermal loading and wall temperatures (iii) Thermal loading changes in direct proportion to the cube root of the mean piston speed. c. For the same intake manifold pressure the wall temperature changes in direct proportion to the thermal loading. 2. For the operation of turbocharged engines at constant power output and no aftercooling, one of the following two conditions prevails: -92

-9a. High supercharging resulting in better fuel economy b. Low supercharging resulting in pooetter fuel economy but b. Low supercharging resulting in poor fuel economy but reduced thermal and mechanical loadings, without any reduction in the wall temperatures. 3. The turbocharged engine performance is improved by increasing the air density and reducing its temperature in the intake manifold resulting in higher specific power outputs and lower thermal loadings and wall temperatures. Aftercooling is one method used for increasing air density by decreasing the air temperature. Even though the pressure drop increases as the aftercooler effectiveness increases, general engine performance will be improved by using an aftercooler with high effectiveness. Improving the compressor efficiency will also increase the specific power output, and reduce the thermal loading and wall temperatures. 4. Aftercooling is necessary if high power outputs are to be obtained at high turbocharging pressures. Recommendations: It is recommended that further investigations be made along the following lines: 1. Apply the equations derived in this work to turbocharged engines operating with aftercoolers and, for tests at high supercharging pressures, compare the measured values for the engine performance, thermal loading, and wall temperatures, with those predicted from these equations.

-942. Study the relationship between aftercooler pressure-drop and effectiveness required for maximum air-density and minimum airtemperature in the intake manifold. These two conditions will give maximum specific power output and minimum thermal loading and wall temperatures. 3. Make a study of possible methods for effectively cooling the air after its compression in the turbocharger. 4. Investigate the effect of valve timing and exhaust manifold pressure and both the cylinder wall temperature and energy of the exhaust gases necessary to drive the turbocharger. 5. Study the distribution of the heat losses during the cycle, over the cylinder of the turbocharged engine.

APPENDIX A SAMPLE CALCULATIONS Run Number 95 Test Conditions: Pb Barometric pressure = 29.52 inches Hg. P Air gauge pressure at the intake manifold in surge tank = 3.1 inches Hg. Pa Air gauge pressure at inlet tap of orifice meter = 3.7 inches Hg. APa Differential pressure across the air orifice = 7.05 inches water. APc Differential pressure across the cooling water orifice = 8.3 inches water. T Air temperature at air orifice = 79 ~F a Tm Air temperature at intake manifold= 90 OF = 550~R T Oil temperature in crank case = 164 ~F o Tw.ex Exhaust manifold wall temperature= 202 ~F TQ Liner wall outer side temperature 1 at T.D.C. = 225 ~F T2 Liner wall outer side temperature 2between T.D.C. and B.D.C. = 201 ~F T Liner wall outer side temperature 3 at B.D.C. = 191 OF T Average temperature of the engine outside wall = 141 ~F ATc = (Tc -Tc ): Rise in water temper3 1 ature 9.95 OF Tc Cooling water temperature at 2 inlet to barrel = 167 ~F TC3 Cooling water temperature at exit from cylinder head = 173 ~F -95

-96Tex Exhaust gas temperature = 875 ~F tF Average time for comsumption of 0.1 lb of fuel = 1.433 minutes Nt Tachometer reading for the time tF= 1176.8 revolutions tF WB Brake load = 42.9 lbs. WF Friction load = 13.06 lbs. N Revolutions per minute 1176.8 821 1.433 2LN 2 x 5.25 x 821 ft S Mean Piston speed =-l= 12 x 60 = 11.98 ft sec H.V. Heating Value of fuel = 19500 B.T.U. lbs. Air Flow Rate: D Actual inside diameter of 2" pipe = 2.0052 inches Al Area of the 2" pipe = v x (2.0052) = 3.16 sq. ins. D2 Diameter of the sharp edged orifice = 0.7018 sq. ins. A_ Area of the sharp edged orifice = x(0.7018)2 = 0.387 sq. ins. Pa Absolute air pressure at inlet tap of orifice meter =29.52 + 3.7 =33.22 inches Hg. p _ 7.05 62.2 lbs. AP = 0.254 a 12 144 sq. in. APa 0.254 Pa 33.22 x 0.49 0.0156 Assume the Reynold's number = 14,300 K (2) Flow coefficient [ for D2 0.7018 = 0.35 ] = 0.6155 (2) D 2.0052 E Area multiplier for thermal expansion of the orifice plate (stainless steel) = 1 Y (2) Emperical expansion factor [ for Pa =.0156] = 0.995 Pa Pa Density of air at inlet to orifice Pa _ 33.22 x 0.49 x 144 = 0.0816 lbRTa 535.534 (460 + 79) cu. ft. References are given in Bibliography. References are given in Bibliography.

(2) Air flow rate = 0.668.A2.K.E.Y. PaPa sbs a... a.a s e-c = 0.668 x 0.387 x 0.6155 x 0.995 o.0o816 x 0.254 = 0.0228 lbs sec Air flow rate = 0.02275 x 3600 = 82.1 lbs hr Check on Reynold's Number: GD Re = 1 _ p. air G = Air flow rate per sq. ft. hr. Pair = Absolute viscosity of air lb ft. hr. Re: 82.1 x 144 x 2.0052 4,3500 3.16 x 12 x 0.0438 Cooling Water Flow Rate: The orifice meter used for measuring water flow rates had the same dimensions as that for air. Using the same symbols used for the air flow rate, the equation for the rates of cooling water flow is )c = 0.668 A2.K.E. /PcAPc s. sec. Assume the Reynold's Number = 19700 K = 0.6123 E at T = 163~F for stainless steel = 1.0011 p at T = 163~F =61 lbs. AP x 62.2 = 0.2986 lb. 12 144 sq. in. A2 = 0.387 sq. ins. Wc - 0.668 x 0.387 x 0.6125 x 1.0011o 61 x 0.2986 = 676 lb. sec. = 0.666 x 3600 = 2435 lb hr

-98Check on Reynold's Number: lb. j water at 163~F = 0.944 ft se Re = 2435 x 144 x 2.0052 19650 3.16 12 x 0.944 Fuel Rate of Consumption: 0.1 0.1 Consumption per hour = - x 60. x 60= 4.19 lbs. t - 1.433 6 = 4.19 lbs. F: Fuel air ratio = 4.12 = 0.051 A 82.1 Power Output: The equation for the engine horse power as measured by the dynamometer is: WxN Horse Power - W 4000ooo where W = load on the dynamometer arm. N = revolutions per minute. 4000 = a constant for the dynamometer. WB xN B.H.P. Brake horse power =WB x 4000 42.9 x 821 = 8.8 H.P. 4000 WI Indicated load = WB + WF = 42.9 + 13.06 = 55.96 lbs. B.M.E.P. Brake mean effective pressure = B.H.P. x 33000 x 2x 12 V N s 3 where Vs is the swept volume = 83.48 in.5 B.M.E.P. (uncorrected) = 8.8 x 00 x 82 x 12 = 101.9 bs. 85.48 821 1sq. in. Correction factor of the B.M.E.P. for the effect of back pressure. = 0.6 - sq. in. B.M.E.P. = 101.9 - 0.6 = 101.3 lbs* sq. in. B.H.P. = 101.3 x 83.48 x 821x = 8.76 33000 2 12

-99n^m Mechanical efficiency WB 42 76.7 WI - 76.7 % 101.3 lbs. I.M.E.P. = = 132.767 sq. in. I.H.P. = 132 x 83.48 821 1 11.42 H.P. 33000 2 12 B.S.F.C. Brake specific fuel consumption 4.19 lb. 8.76 = 477 B.H.P. hr. I.S.F.C. = 4.19 = 0.66 lb 11.42 I.H.P. hr. 1B.Th. = Brake thermal efficiency 8.76 x 550 x 60 x 60 x 100 o = 27.5 778 x 4.19 x 19,500'I.ITh. = Indicated thermal efficiency B.Th 27.55 5.62 %9 ~m ^0.767 Heat to cooling water = Wc x sp.ht. x ATc B.T.U. = 2435 x 1 x 9.95 = 24220 hr. (P-V) and (Wall Temperature-Crank Angles) diagrams: Scale of pressure Calibration pressure = 231.3 lb sq. in. This corresponds to a calibrating resistance of 1 ohm 2 in the "Bridge-Amplifier' Deflection on the oscillograph screen, figure (23) = 4.5 screen divisions. Scale of trace = 231.3 =51.5 lbs. per division. 4.5 sq. in. Scale of temperature Voltage across the thermocouple ends (as measured by the precision potentiometer) = 0.585 m.v.

-100Voltage corresponding to 100~F temperature difference between thermocouple junction and reference junction temperatures (In the range of 40~C to 150~C, figure (18) = 1.765 m.v..585 Temperature difference corresponding to.585 m.v.=' x 100 = 33.1~F Corresponding deflection on the oscillograph screen = 7 screen division Scale of trace = 31 =4.73 -F 7 division Weight of the Residual Gases in the Clearance Volume: Assume the pressure at the end of the exhaust stroke = 14.7 lbs. sq. in Clearance volume = 6.393 cu.ins. R: Gas constant for exhaust gases (20) = 53.175 ft. lb. ex lb. ~F. Wt.of residual gases = PV = 14.7 x 144 x 6.393 =.000109 lb. per cycle RT 53.175 x 144 x 12 Wt. of scavenging air per sec. =.00764 x 57 = 0.00526 lb (Appendix F) 16.3 550 sec Time for scavenging per cycle = 60 x 56 = 0.00731 sec. 821 x 560 Wt. of scavenging air per cycle = 0.00526 x 0.00731 = 0.0000385 lb. Wt. of fresh air trapped in the cylinder per cycle 82.1 x 2 =- o0. 0000385 60 x 821 = 0.003,33 - 0.000,039 = 0.00329 lb. Wt. of gases during the compression stroke = 0.00329 + 0.00011 = 0.0034 lb. 4.19 =. Wt. of fuel used per cycle = 1 = 0.00017 lb. g21 x 60 Wt. of gases during the expansion stroke = 0.0034 + 0.00017 = 0.00357 Ib.

-1011000 IiII I I I I I I 900 800- 0o~ 0 o 700 - 0 600 500. 400 - 300 I 0 200 - 100 - 90 - 80 so 70 - 60 50 - Volume cuNin9 40 2 3 4 5 6 7 8 9 10 20 30 40 50 60 70 so 90100 630F -51 20U Volume cu.in. I 2 3 4 5 6 7 8 9 10 20 30 40 50 60 708090100 P-V DIAGRAM (Run Number 95) FIG. 51

-102For any point on the compression stroke, T Px V = o.458 P.v. 12 x 53.34 x 0.0034 - For any point on the expansion stroke, T = x V= 0.438 P.V. 12 x 535.175 x 0.00357 To draw the pressure-crank angles diagram near B.D.C., allow for all the runs, 0.75 lb.per.sq.in.drop in air pressure below the manifold pressure at the time the I.V. closes..'. Pressure at close of I.V. = 32.62 x 0.49 - 0.75 = 15.27 lb. sq. in. Allow, for all the runs, 20~F drop in gas temperature between the end of the expansion stroke and the measured exhaust temperature,. T at the end of expansion stroke = 875 + 20 = 895~F = 155~R & P at the end of expansion stroke = 1355 = 34.4 lbs. 0.458 x 89.87 sq. in.

-103Heat Transfer Analysis For The Cycle: Equation [2.1]: ag = 0.0564 3J PT for this run 2iLN 2 x 525 821 1 ft. S x 11.98 ft. 60 - 12 60 sec. thus ag at any crank angle where the gas is at a pressure P and temperature T = 0.0564 411.98 JPT = 0.1291 PT The values of COg and (ag.Tg) for the compression and expansion strokes are shown in tables (1) and (2). For the intake and exhaust strokes, the method used is as follows: Intake stroke: For the intake stroke the effective gas temperature was calculated by equating the heat gained by air to the heat transferred from the walls. oa x cp x At = Am x eg x (Twg- TE) [10.1] where At is the temperature rise in the mixture of air and residual gases, due to the heat from the walls. The temperature of the mixture of residual exhaust gases and air was calculated from the following equation: (Wair + %exh.) mix. = Wa x ha + xexh. x hxh. B.T.U. ha at 90~F = 131.46 B. hexh. at 8750F = 321.5 B.T.U. lb.

.00333 x 131.46 +.00011 x 521.5 D B.T.U. h m. = —----------- 137 4 l.w. mix..00333 +.00011 b. Tmix. =115F Temperature of the gases at the time when I.V. closed PV 15.27 x 144 x 82.97 582R 122F oDR 144 x 12 x 0.0034 x 53.34 Rise in gas temperature due to heat from walls 122 - 115 = 7~F 75% of this heat was assumed to be gained during the intake stroke, and the rest during the compression stroke before the inlet valve closed. Heat gained during the intake stroke = 82.1 x 0.24 x 7 x.75 = 103.5 B.T.U. hr. Mean wall temperature during the suction stroke = 212~F (Appendix E) To get the coefficient of heat transfer between the gases and wall, assume 1 lb. per sq. in. drop in air pressure below the manifold pressure during the intake stroke. lbs.. P = 32.62 x.49 - 1 = 15.02 bs sq. in. Average gas temperature = 122 + 115 =118.5~F = 578.5 ~R aM = 0.1291 415.02 x 578.5 12.02 B.T.U. hr. sq. ft. ~F AM = Combn ch. area + 1 liner area + intake manifold area = 48.4 + 42 + 34.3 = 119.8 sq. ins. From equation [10.1] Tw. -TE = 103.5 x 144 = 10.35~F.g. 12.02 x 1 9.8

-105TE = 212 - 10.35 = 201.65~F = 661.65~R xMTE = 12.02 x 661.65 = 0.796 x 104 Exhaust stroke: The gas pressure during this stroke is assumed to be 1 lb. higher than the atmosphic pressure = 14.5 + 1 =15.5 lbs. sq. in. sq.in. Exhaust gas temperature = 875 + 460 = 1335~R D.T.U. Mc = 0.1291 15.5 x 135 = 18.56 hr. sq. ft. - F cMT =18.56 x 1335 = 2.48 x 104 At the end of the expansion stroke the pressure was considered half way between the exhaust pressure and that from extrapolation of the expansion curve. The temperature was found by extrapolation. Values Of aM And TM.E As Calculated From Figures 52 and 53: Stroke Intake Compression Expansion Exhaust Total Cycle CM 12.02 26.55 70.5 18.56 31.86'MTM.E. 0.796 x 104 2.415 x 104 15.5 x 104 2.48 x 104 5.298 x 104 TM.E.oR 661.7 910 2205 1335 1660 TM.E.OF 201.7 450 1745 875 1200

-106TABLE I Heat Transfer Analysis For The Cycle: (Run #95) Crank Deflection Gauge P V T Stroke Angles of trace Pressure Absolute Volume Temper- 1WP a Cg T From Pressure ature g T.D.C. screen lbs. lbs. x 10 divisions sq. in. sq. in. cu. ins. ~R 0 10.2 520 534.5 6.393 1552 910 117.6 18.25 3 8.3 428 442.5 6.474 1305 760 98.1 12.8 6 7.3 376 390.5 6.675 1186 68o 87.8 10.4 9 6.8 350 364.5 7.031 1166 652 84.2 9.82 12 6.2 319 333.5 7.613 1157 620 80.1 9.27 15 5.45 281 295.5 8.158 1098 570 73.6 8.08 18 4.8 247 261.5 8.946 1063 528 68.2 7.26 21 3.9 201 215.5 9.6 942 452 58.4 5.5 107J H 24 3.25 167. 181.8 10.887 899 404 52.2 4.69 ~ 27 2.7 139 153.5 11.998 837 358 46.3 3.88 ou 36 - - 108 16.175 794 293 37.85 3.03 45 - - 78 21.22 754 242.5 31.35 2.36 63 - - 45 33.3 682 175 22.6 1.54 72 - - 36.6 39.95 666 156 20.16 1.34 90 - - 25.8 53.32 626 127 16.4 1.03 108 - - 20.2 65.72 604 110.4 14.28 0.86 126 - j - 16.8 76.05 581 98.8 12.76 0.74 142 - - 15.25 82.97 575 93.7 12.11 0.7 180 -. - 15.02 89.87 573 92.8 12 0.69

-107TABLE II Heat Transfer Analysis For The Cycle: (Run #95)'Crank Deflect i, Gauge P V T Stroke Angles of trace Pressure Absolute Volume Temper- PE ag.gTg From Pressure ature T.D.C. screen Ibbs. lbs. 10-4 division sq. in. sq. in. cu.ins. ~R 0 10.2 520 534.5 6.393 1552 910 117.6 18.25 3 12.7 654 668.5 6.474 1890 1123 145.2 27.42 6 15.55 800 814.5 6.675 2377 1390c 179.6 42.6 9 16.2 835 849.5 7.031 2610 1490 192.5 50.3 12 15.5 798 812.5 7.613 2706 1482 191.4 51.9 18 13.3 685 699.5 8.946 2740 1383 178.8 49 21 12.25 631 645.5 9.60 2710 1320 170.5 46.2 27 9.8 505 519.5 11.998 2720 118 153.3 41.7 36 6.9 355 369.5 16.175 2612 982 127 33.17 45 4.9 252 266.5 21.22 2475 812 105 26 o Ci 48 4.15 214 228.5 23.103 - - 51 3.6 185.5 200 24.99 - - - 54. 3 170 184.5 27.013 2180 634 81.6 17.8 57 2.95 152 166.5 29.073 - - - 60 2.6 134 148.5 31.163 - - - 63 - - 137 33.3 1992 522 67.5 13.45 72 - - 107 39-95 1868 447 57.8 10.8 90 - - 72.5 53.32 1690 350 45.2 7.65 108 - - 55 65.72 1580 294.2 38 6 126 - - 45 76.05 1497 259.5 33.5 5.01 135 - - 42 80.24 1475 248.5 32.15 4.74 180 - - 25 89.87 1353 184 23.75 3.22

4 -p C,-j CoefIficient of heat 1ransfer From Gas -to Wl3z14 For ihe Cycle 200 OF ILuji Ho. 95 180 160 140 120_ \8 100 80 60 360 315 270 225 1 s 135 90 45 0 45 90 135 CSO 225 270 315 360 T.D.C. B.D.C. T. D. C. B.D.C. T.D.C. Crank I Li-'Clake Compression Expansion Exhaust FIG. 52

+60_ E C(GI(. (c.Tg) For the Cycle " ~m I ~ ~ of Run Hio. 95 50 40 30 20 10 360 315 270 225 180 135 90 45 0 45 90 135 180 225 270 315 360 T.D.C. B.D.C. Tr.D.C. T.D.C. Intake Compre s sion Expana lion Exhaust FIG. 53 Cralk Anigles

a 130 12cd 120_ SURFACE AREA VARIATIOiT < \ L \ VS. 1/\10_ / CRANK ANGLES 100 90807060 010~ ~ ~~~~~ 0 FIG. 54 ~40_Q. p 30_ 200 H 10- ^ I I II 1 L II I 360 315 270 225 180 135 90 45 0 45 90 135 180 225 270 315 360 T.C 13 C T.D CC. C. T.D.C. B.D.C. T.D.C, Crak An Intake Co(ipre s L ion Expansion EiausFG t FIG. 54

APPENDIX D CALCULATION OF CONSTANT "C" OF EQUATION [2.11] Run 95: Area of liner in zontact with water = 133.3 sq. ins. Area of combustion chamber in contact with water = 29.0 sq. ins. Area of exhaust manifold in contact with water = 46.9 sq. ins. Total area = 209.2 sq. ins. Heat carried by cooling water (table 5) = 24220 1T.U. hr. Average wall temperature figure (55) = 199 ~F Average water temperature = 168 ~F (Twall av.-Twater av.) = 199-167.98 = 31 ~F = 24220 x 144 540.T.U. 54 O 209.2 x 31 hr. sq. ft. ~F Ib. i water at the average temperature 0.905. s. ft. sec. (k) water at the average temperature = 2.4 (Wc)06 = 115 c()0.4 = 1.42 k (WC)0.6 x o(Z)0.4 163.2 Similar calculations were made for other runs at various speeds, these are shown in tables (14) and (15). The value Jf the constant C was calculated from the slope of the line in figure (57) and was found to be equal to 3.15 ac- =.15 (Wc )06 (g)0.4 -111

-112Ou. 240' 230 = 220 ~ 210 200 Average Temperature r () 190 180 170 160 - i i V FIG.:: 15 0 -',.. C0 0 0 0 140 o f - 0 O u) 0 0C U 0: 0 20 40 60 s0 100 120 140 160 180 200 220 Liner Comb. Exh. Manifold Area. Sqtins. CALCULATION OF THE AVERAGE WALL TMPERATURE I I FOR RUN N0, 95 FIG. 55

APPENDIX C CALCULATION OF THE COMBUSTION CHAMBER WALL TEMPERATURE FOR RUN NO. 95 1. Using equation [2.14] for the combustion chamber temperature Tw.g. = TM.E. [ Z + Tc z + kw +kw ~M ~M where Twhere T Tm 14.71 + 0.233 I.M.E.P. *ME* 3'TmP 0.48 + 0.00157 I.M.E.P. = 1750 R = 1290~F (Table 16) M B 2.7',TT.U.' (Table 16) hr. sq. ft. ~F Aw g. Aw.g. kw Am + Aw.c. c x = 0.282 in. am: Aw- = 0.877 Am a: ** = 0.782 C Aw. c. k = 27 D.T.U. w hr. sq. ft.(F) CX = 3.15 (Wc)O-6 x (i)0.4 jL k i and k were evaluated at T3, a = 35.15 ( )06 x (2.3)0 4 ~~= 515.B.T.U. hr. sq. ft. ~F Z** = 0.738 Tc was evaluated at exit from engine = 633 ~R. ~ Tw.g. = 709.3 ~R = 249.3 ~F To be compared with Tw. g measured from figure 22, which equals = 243.6~F % error = 2.34 % -113

To draw figure (56) for the imaginary walls on the gas and coolant sides of the wall, the following was calculated for run number 95: Effective combustion chamber wall thickness Aw.g. x = 0.282 x 0.877 = 0.247 in. Am Thickness of the imaginary wall on the gas side = k = 27 x12 9.91 in. cM 32.7 Thickness of the imaginary wall on the cooling water side = kw. g. = 27 x12 x 0.782 = 0.491 in. ac Aw.c. 515 2. Piston Crown Maximum Temperature: Piston crown thickness = 0.75 ins. B.T.U. k: thermal conductivuty = 27 hr sq. ft. ( hr. sq. ft. F L) ft aa = o.o564 5J PT where P = 14.7 T (oil in crank case) = 164~F = 624~R S = 11.97 ft. sec. B.T.U. 12.38 *ua = 128 hr. sq. ft. ~F Thickness of the imaginary wall on the crank case side = k 27 x 12 26.2 ins. aa 12.38 Figure (56) drawn with the calculated imaginary walls thickness, shows that the pistin-crown maximum temperature equals 990~F compared to 250~F for the combustion chamber walls.

13 00 TME. 1200> \ N FPISTON-CROWN THICKNESS 0.75 in. A~\ ^, N~ EIMAGINARY WALL ON CRANK1100 \ CASE SIDE 26.2 in. PISTON TEMPERA1000- \ TURE 990OF 900 0~ 800 CALCULATION OF \700 \ Tw. g. - COMBUSTION CHAMBER WALL TEMPERATURE |g~~~~~~~~700 — l~ \~ ITp.g. - PISTON-CROWN MAXIMUM TEMPERATURE 600 500 COMBUSTION CHAMBER 400 WALL 300 COMBUSTION CHAMBER TEMPERATURE 250~F 200 Tc " 100 IMAGINARY WALL ON COOLING WATER SIDE 0.491 in. IMAGINARY WALL ON GAS SIDE 9.91 in. COMBUSTION CHAMBER WALL THICKNESS 0.247 in. Figure 56

APPENDIX D CALCULATION OF INTENSITY OF THERMAL LOAD ON THE COMBUSTION CHAMBER WALLS. For Run Number 95: Using equation [2.16] 1 1 x 1 - = + am x + ac X U a x k+, a, 1 -+.877 x.282 +.782 x 1 32.7 27 x 12 515 -.0529 U = 50.4 B.T.U. hr. sq. ft. ~F Substituting in equation [2.15] q = U (TM.E.-T,) = 30.4 (1290 - 170) = 34050 B.T.U. hr. sq. ft. -1.16

-117TABLE III Engine Test Results - Series A Runs at Pm 30" Hg, Tm % 80~F, N % 800 R.P.M. Run Number 1 2 3 4 15 85 94 147 148 Pb ins. Hg. 29.15 29.18 28.96 28.96 29.42 29.36 29.52 28.96 28.96 Pm ins. Hg. 28.75 28.78 28.56 28.56 29.02 29.01 29.55 28.56 28.51 Tm OF 83 84 72 72 77.5 82 90 76 76 N R.P.M. 807 800 800 810 854 775 826 803 823 To *F 145 149 145 141 143 153 164 146.5 147 Tw.exh. F 171 197' 191.5 200 167 187 193 164 157 Tl ~F 184 207 201 207 181 206 219 177 169 T2 *F 173 188 186 188 173 192 199 168 171 T3, F 168 180 178 179 170.5 185 188 165 166 T8 ~F 126 118 115 114 117.5 132 140 124 127 ATC - (TC3-TC) 5.76 7.88 6.845 8.82 5.27 7.4 8.8 4.32 5.78 TC2 F 142.86 - 153.4 - - 156 162 - T *F 146.2 161.7 158 161.6 163 162.7 169 155.5 151 Texh. OF 438 747 658 944 368 603 856 348 307 Wa lbs./hr. 70.6 69.7 69.7 69.3 75.6 69 69.6 71.8 73.8 Wc lbs./hr. 2412 2400 2400 2424 2540 2322 2460 2400 2520 WF lbs./hr. 1.843 3.22 2.668 3.66 1.471 2.38 3.59 1.363 1.154 F/A Ratio.0261.0463.0383.0528.0195.0345.0515.019.01566 B.H.P. 2.75 6.34 5.14 7.52 1.7 4.29 7.31 1.332.729 B.M.E.P. 32.35 75.3 61 88 18.9 52.3 84 15.8 8.4 Imech. % 52.9 72.5 68.6 75.7 39.2 64.6 74.6 35.1 22.1 lbs. I.M.E.P. lbs. 61.2 101 88.9 116.2 48.3 81 112.5 45 38 B.S.F.C. lblh-.671.508.52.487.867.555 0.491 1.023 1.584 B.H.P. hr. I.S.F.C. lbs..355.368.357.369.34.358.366.359.35 I.H.P. hr.'1B.Th. % 19.45 25.7 25.1 26.8 15.05 23.52 26.58 12.75 8.24 1I.Th. % 36.77 35.5 36.6 35.4 38.4 36.4 35.6 36.4 37.25 Qcooling B.T.U. 13900 18900 16410 22400 13400 17200 21650 10370 9520 hr.

TABLE IV. Engine Test Results - Series A Runs at Pm 335" g, Tm 80~F, N~ 800 R.P.M. Run Number 30 32 5533 36 37 81 93 95 146 149 Pb ins. Hg. 29.2 29.2 29.2 29.2 29.28 29.39 29.52 29.52 28.96 28.96 Pm ins. Hg. 32.45 32.24 52.7 32.3 31.93 32.84 52.55 32.62 32.01 32.26 Tm ~F 75.5 79 75.5 87 77 86 87.5 90 76.5 75 N R.P.M. 793 855 809 790 830 796 832 821 818 816 To ~F 151 157 157 155 139 150.5 160 164 142 Tw.exh. F 174.5 184 187.5 195 206 167 186 202 159.5 i56 TAl ~F 189 202.5 203 206.5 216 183 208 225 175 168 TB2 ~F 181 190.5 190 190.5 196 177.5 191 201 168 164 TA3 ~F 176.5 182 182.5 182.5 184 172 181 191 165 160 Ts OF 120.5 122.5 125 123 111 132 139 141 124 126 Tc = (Tc -TC) 6.74 7.6 8.68 9.0 10.62 6.5 8.75 9.95 4.15 3.78 T ~F - 160 162 154.5 156 154 158 167 - C2 T O~F 165.7 164.6 167.7 165 163 160 166 173 154.2 149.8 Texh. ~F 418 592 659 759 962 574 755 875 322 280 W lbs. 78.5 84.5 81 78.4 80.8 81.6 83.5 82.1 83.3 83.25 W Ibs. 2368 2550 2418 2365 2475 2382 2478 2435 2430 2430 hr. Wfuel lbs. 1.97 2.982 5.255 3.475 4.41 1.672 3.79 4.19 1.366 1.118 hr. F Ratio.0.0253.0399.0444.0546.0205.0455.0511.0164.01342 A B.H.P. 3.26 5.75 6.48 7.52 9.65 2.14 7.64 8.8 1.456.864 B.M.E.P. 59.05 63.6 76.2 90.5 110.3 25.55 87.1 101.9 16.9 10.12 B.M.E.P. Corr. 57.65 62.5 75.25 89.7 109.8 24 86.5 101.3 15.26 8.39 Tmech. % 55.8 67 71.2 74.5 78 45.4 73.7 76.7 36.65 25.5 I.M.E.P. 67.5 93.4 105.6 120.4 140.8 52.9 117 132 41.7 32.85 lbs. B.S.F.C. B.P. hr..65.551.506.466.46.795.501.477 1.04 1.564 I.S.F.C. lbs..3515.356.3604.347.359.361.369.3663 0.381.399 I.H.P. hr. nB.Th. % 20.7 24.6 25.8 28 28.35 16.4 26 27.35 12.56 8.35 TI.Th. % 37.15 36.62 36.2 37.6 36.35 36.13 35.35 35.62 32.31 32.7 Qcooling B.TU. 15920 19380 20970 21300 26280 15000 21650 24220 10100 9180 hr.

TABLE V. Engine Test Results - Series A Runs at P. 56" Hg, Tm 80~F, NR 800 R.P.M. Run Number 47 48 49 50 51 52 82 87 92 96 Pb ins. Hg. 29.52 29.52 29.52 29.52 29.52 29.52 29.59 29.56 29.52 29.52 PM ins. Hg. 35.12 55.12 55.27 55.42 35.5 55.17 55.29 56.06 55.52 55.92 Tm OF 75 75 77 79 80.5 81 87 85.5 84 88 N R.P.M. 828 841 829 824 850 847 825 751 823 840 To OF 159 149 155 159.5 161 160 156 161 156 166 Teh oF 161 178 185 189 194 207.5 158 178 185 202 Til OF 175 194 203 207 217 228.5 177 201 208 225 Ti2 OF 170 185 189 192 194 203 169 187 189 202 2 T'0 OF 166 179 183 184 187 192 165.5 181.5 181 191.5 Ts OF 108 116 118 121 121 123 134.5 159.5 136 145 ATc= (Tc -Tc ) 5.02 7.0 7.65 8.65 9.7 11.0 6.14 7.94 9.2 10.04 Tc2 OF 149 156 159 160.5 l6o.5 161 145 158.5 158 166 T OF 155 162 166 168 169.5 169 149.5 165 165.5 175 Texh. OF 527 473 550 665 780 905 342 466 724 846 We lbs. 93.7 92.5 91.7 91.8 90.2 90.3 91.1 85.6 90.8 89.8 hr. W Ibs. 2478 2508 2478 2460 2558 2526 2460 2250 2454 2508 hr. Wfuel Ibs. 1.68 2.487 2.945 5.442 3.89 4.685 1.76 2.38 3.65 4.265 hr. F Ratio.0179.0269.0321.0375.0431.0519.0193.0278.0403.0485 A B.H.P. 2.265 4.68 6.2 7.65 9.07 10.46 2.507 4.45 8.56 9.95 B.M.E.P. 26.0 52.9 71.1 88.2 101.5 117.5 28.85 55.6 94 114.5 B.M.E.P. Corr. 24.4 51.6 69.9 87.2 100.7 116.7 27.55 54.5 95 115.8 Thnech.% 47.6 64.6 71.4 75.5 78 80.2 50.4 67.2 76 80 I.M.E.P. 51.5 79.9 98 115.5 129 145.7 54.4 80.8 122.4 142.2 B.S.F.C. lbs.79.545.484.455.431.45.742.554.441.433 B.H.P. hr. I.S.F.C. lbs.-.376.357.345.343.336.361.374.372.35.546 I.H.P. hr. qB.Th. % 16.55 23.95 26.97 28.7 30.3 29 17.6 23.6 29.6 30.17 8I.Th. % 54.7 57.1 57.8 58 58.85 56.17 54.92 55.1 38.9 57.7 Qcooling IB..U. 12420 17580 18940 21200 24600 27800 15100 17850 22600 25200

TABLE VI Engine Test Results - Series A Runs at Pm39" Hg, Tm, 80~F, N 800 R.P.M. Run Number 58 60 61 62 63 64 71 83 91 97 Pb ins. Hg. 29.69 29.69 29.69 29.69 29.69 29.69 29.52 29.39 29.52 29.52 Pm ins. Hg. 58.59 38.59 58.44 39.59 38.99 59.24 39.27 38.54 38.77 38.47 Tm ~F 74.5 76 78 79.5 81 32 90 87 80 83 N R.P.M. 840 864 845 817 827 807 789 820 826 814 To ~F 147.5 155 158 159 159.5 161 159.5 155 157 160 Tw.exh. F 163 178 180.5 186 192.5 199.5 200 174.5 184 198 TI ~F 177.5 196 200 206 214 220 225 191 208 220 T,2 ~F'172 186 188 191.5 196 200.5 200 182 191 202 T~ OF 168.5 178 181.5 185 187 189.5 189.5 179 184.5 192 T ~F 124 129.5 154 155.5 158.5 158 158 157 155 140 ATc=(T c -Te ) 4.55 7.2 7.99 8.46 10 10.99 12.17 5.02 8.67 9.96 T2 ~F 156 160 161 163 163.5 165.5 161 161 158.6 163.5 1 T3 ~F 158.6 164.7 166 167.5 169 171.5 170 166 166.5 170 Texh. ~F 500 468 521 554 704 792 825.5 528 652 724 Wa lbs 101.5 1035. 101.2 98 98.7 99 98 99.9 97.4 95.7 W lbs. 2502 2574 2490 2442 2472 2412 2564 2448 2460 2450 hr. Wuel bs. 1.607 2.62 5.013 5.11 3.91 4.2 4.59 1.74 5.29 3.94 hr. F Ratio.01585.02555.0298.0518.0596.0424.0448.0174.0538.0412 A B.H.P. 2.4 5.25 6.46 7.06 9.22 10.22 10.4 2.66 7.95 9.09 B.M.E.P. 27.15 57.5 75.5 82 105.6 120.4 123.3 50.8 91.5 106 B.M.E.P. Corr. 25.45 56.1 72.5 80.8 104.6 119.5 122.5 29.2 90.2 105.1 imech. % 49.5 67.8 72.4 75 79.4 81.6 82.5 54.9 76.8 79.5 I.M.E.P. 51.5 82.8 101 107.8 152 146.5 149 55.3 117.5 152.5 lbs. B.S.F.C. B..P. hr..714.514.4695.448.43.4135.452.69.42.438 I.S.F.C. lbs..553.548.5595.556.5412.557.355.579.5225.548 I.H.P. hr. lB.Th. % 18.3 25.4 27.8 29.15 50.55 31.6 30.2 18.91 51.1 29.77 nI.Th. % 36.95 57.5 58.45 58.85 58.2 38.7 36.7 54.46 40.5 37.45 Qcooling B.T.U. 11400 18500 19900 20670 24720 26500 28800 12500 21350 24200 hr.

TABLE VII. Engine Test Results - Series A Runs at Pm 42" Hg, Tmt80~F, N%800 R.P.M. Run Number 66 67 68 69 70 84 89 90 Pb ins. Hg. 29.52 29.52 29.52 29.52 29.52 29.56 29.36 29.52 PM ins. Hg. 41.47 41.97 41.47 41.72 41.97 41.26 41.71 41.8 Tm OF 76 82 85 86.5 88.5 86 85 72.5 N R.P.M. 827 786 831 805 815 805 775 837 To "F 149 154 158 158 159 160 159 134 Tw.exh OF 168 178 181 186 198 164 176 182 Teg OF 181.5 195 202 206 220.5 179 200 206 T2g ~F 175 184 188 190.5 200 172 185.5 190 T 3 OF 172 179.5 180.5 183 190 168 180.5 181 Ts OF 128 134 136 138 138 135 137 126 ATC(T= c-T1 )5.185 7.9 8.51 9.95 12.83 5.1 7.9 8.52 T2 "F 157.5 158.5 158 159 159.5 150 158.5 157 T OF 160 162.5 164 165 169 155 164 164 H Texh. OF 303 416 535.5 626 799 305 397 610 Wa lbs. 109 106.8 109.8 105.8 106.2 106.5 100.3 106.2 hr. WC Ia. 2466 2358 2484 2412 2430 2412 2322 2502 hr. Wfuel lbs. 1.674 2.551 3.26 3.7 4.61 1.69 2.485 3.565 hr. F Ratio.01535.0239.02975.035.0434.0159.0248.0336 A B.H.P. 2.84 5.18 7.75 8.75 11 2.89 4.84 8.08 B.M.E.P. 32.05 62.5 88 102.7 127.2 34.2 59.3 91.7 B.M.E.P. Corr. 30.35 61.1 86.7 o101.6 126.4 32.6 57.9 90.5 rmech. % 55.4 71.2 77 80 83.2 57 69.8 78 I.M.E.P. 54.8 86 112.6 127 152 57.3 83 116 lbs. B.S.F.C. B.H.P. hr..634.505.43.43.425.611.526.447 I.S.F.C. ljbs. I.S.F.C...351.3595.331.344.354.348.367.3485 I.H.P. hr. IB.Th. % 20.6 25.83 30.4 30.36 30.72 21.38 24.8 29.2 I.Th. % 37.2 36.3 39.5 37.92 36.9 37.45 35.55 37.4 &cooling B.T.U. 12800 18600 21170 24000 29550 12300 18350 21200 hr.

-122TABLE VIII. Engine Test Results - Series A Runs at Pm 45" Hg, Tm 80~F, N 800 R.P.M. Run Number 72 73 74 75 76 Pb ins. Hg. 29.45 29.45 29.45 29.45 29.45 P ins. Hg. 44.25 44.25 44.05 44.55 44.7 m Tm ~F 83.5 85 88 89.5 90 N R.P.M. 823 829 832 815 813 TO ~F 151 154.5 158.5 160 161 Tw.exh. F 159.5 179 178.5 184 193.5 Ti1 ~F 175 199.5 201 208 216.5 T12 ~F 169 185 187 191 198 T3 ~F 166 181 181 184 189.5 T ~F 128.5 133 138 137 138 s \TC= (TC -TC1) 5.19 5.4 9.19 9.21 10.5 T ~F 146 - 156 160 163 c2 T ~F 149 162.5 163 167 170.5 C3 Texh. F 290.5 386 495 532 668 W lbs. 116.3 116 117.2 112.5 110.8 a hr. WC lbs. 2454 2472 2484 2430 2430 hr. Wfuel lbs. 1.788 2.427 3.01 3.38 3.96 hr. F Ratio.01538.02095.02565.0300.0357 A B.H.P. 3.04 4.94 6.78 8.13 9.45 B.M.E.P. 35.1 56.6 77.5 94.6 110.5 B.M.E.P. Corr. 33.4 55,1 76.2 93.4 109.5 Vmech. % 57.6 68.6 75.2 78.6 81.1 I.M.E.P. 58 80.4 98.6 118.6 135.2 B.S.F.C. lbs..618.505.451 422.419 B.H.P. hr. I.S.F.C. lbs..356.3465.339.3317.34 I.H.P. hr. nB.Th. % 21.12 25.8 28.9 30.92 31.2 WI.Th. % 56.65 37.62 38.45 39.3 38.45 Qcooling B.hr. 12720 13350 22800 22400 25500

TABLE IX. Engine Test Results - Series B Runs at Pmv 36" Hg, Tm 80~F, N S 1200 R.P.M. Run Number 99 104 121 129 130 134 135 136 137 Pb ins. Hg. 29.25 29.25 29.16 29.03 29.03 29.20 29.20 29.08 29.08 PM ins. Hg. 35.275 35.45 35.225 55.83 35.53 35.45 35.65 35.48 35.55 Tm ~F 82.5 90.5 84 80 87 81 83 82.5 86 N R.P.M. 1240 1197 1210 1208 1190 1227 1216 1223 1212 To ~F 172.5 177.5 171 167.5 174 159 165 162 169.5 Twexh. "F 181 194.5 176 170 192 172 175 177 187 Tl O~F 201 218 197 188 216 194.5 196 200 207 T12 ~F 189.5 200 184 178 190 179 182 179 188 Te3 ~F 182 191 178 175 179 173 176 173 179 Ts ~F 145 149.5 142 141 146 134 140 136 142 &Tc=(Tc3-Tc1) 6.05 8.17 6.26 5.49 9.19 6.35 6.6 7.69 7.72 T2 ~F 160 163 - - 149.5 - - - 153 T3 "~F 165.5 169.5 157 159 158 157 159 155 159 - Texh. OF 534 772 578 403 972 588 614 810 838 8O Exhaust Back Press. -- - -- 29.25 34.20 29.23 34.15 ins. Hg. Wa Hs 137 133.75 134.5 135.2 133 135.5 136 135.5 132.8 Wc lbs. 3642 3032 3545 3560 3510 3600 3565 3600 3550 hr. W uel bs. 3.39 4.84 3.64 2.51 5.99 3.738 3.808 5.15 5.1 fuel hr hr. F Ratio.0246.0362.02905.01855.0451.02759.028.038.0385 A B.H.P. 6.61 11.49 7.92 4.05 14.37 7.93 7.69 11.85 11.6 B.M.E.P. 50.9 91.2 62.3 31.8 114.8 61.4 60.1 92 91 B.M.E.P. Corr. 49.5 90.15 61.05 30.2 114 60.1 58.8 91 90 nmech. % 60.8 73.9 65.7 49.5 78.2 65.5 63.15 73.9 72.8 I.M.E.P. 81.5 122 92.9 61 145.9 92.2 93.2 123.1 123.6 lbs. B.S.F.C...P. hr..528.426.4679.654.42.481.508.439.445 I.S.F.C. bH.sPhr..3205.314.3078.3238.3281.314.3205.3242.3238 nB.Th. % 24.8 30.02 27.81 21 18.6 27.15 25.78 29.78 29.4 II.Th. % 40.8 40.7 42.4 42.5 23.8 41.6 40.85 40.39 40.3 Qcooling Bhr. 22000 24750 22200 19550 32250 22820 23550 27620 27400

-124TABLE X. Engine Test Results - Series B. Runs at Pmt36" Hg, Tm P 140~F, N P 1200 R.P.M. Run Number 122 126 127 128 131 Pb ins. Hg. 29.27 29.42 29.42 29.28 29.03 Pm ins. Hg. 35.63 35.72 35.845 35.78 35.56 Tm ~F 137 142 145 144 141 N R.P.M. 1201 1216 1216 1204 1204 To ~F 176 173 175 173 178 Tw. exh. OF 182 183 184 175 199 Tg i ~F 208 208 211 199 224 T02 ~F 191 192 188 187 200 T 3 ~F 186 184 178 180 190 Ts "F 145 144 145 144 148 ATC = (T 5-Tcl) 6.64 7.05 8.52 6.35 9.08 T2 ~F- 158 - 156 160 T ~F 166 164 157 162 168 c3 Texh. OF 640 697 865 514 920 Wa lbs. 125.1 126.7 125.2 126.2 122.4 a hr. Wc lbs. 3535 3585 3585 3550 3550 hr. Wfuel lb. 3.76 3.94 4.97 3.01 5.735 F Ratio.0301.0311.0397.0239.0468 A B.H.P. 7.78 8.54 11.55 5.125 12.58 B.M.E.P. 61.6 66.75 90.4 40.4 99.2 B.M.E.P. Corr. 60.4 65.55 89.45 55.9 75.9'mech. % 65.6 67.8 74 55.9 75.9 I.M.E.P. 92 96.7 120.8 70 129.8 B.S.F.C. lbs. G B.S.F.C. hr.495.471.4572.6075.459 I.S.F.C. lbs..3252.319.3215.539.348 I.H.P. hr. 1B.Th. % 26.5 27.75 30 21.5 15.57 TI.Th. % 40.4 40.9 40.6 38.5 20.55 cooling B.T.hr. 23450 25500 50580 22550 52200

-125TABLE XI. Engine Test Results - Series B Runs at Pm% 36" Hg, Tm,200~F, N P 1200 R.P.M. Run Number 123 124 125 132 133 Pb ins. Hg. 29.27 29.27 29.27 29.12 29.12 Pm ins. Hg. 35.65 35.49 35.70 35.75 35.62 Tm ~F 193 198 198 197 204 N R.P.M. 1212 1231 1212 1211 1214 To ~F 177 176 175 178 178 TwV exh. OF 186 200 187 174 184 T2 ~F 214 230 213 199 213 T12 ~F 195 206 196 186 190 T13 ~F 188 195 187 180 180 T, ~F 148 148 146 148 150 AT=' (Tc3-Tc1) 7.13 9.02 7.25 6.19 7.97 T F - - - - 154 T ~F 168 173 165.5 163 159.5 C Texh. OF 688 949 700 475 807 W lbs. 110 109 109.2 119.9 118.5 a ihr. WC lbs. 3565 3617 3565 3558 3580 hr. Wfuel lbs. 3.675 4.98 3.69 2.555 4.355 hr. E~ Ratio.0334.0457.0338.0213.0367 A B.H.P. 7.68 11.62 7.83 3.755 9.76 B.M.E.P. 60.2 89.6 61.25 29.42 76.25 B.M.E.P. Corr. 59.1 88.8 60.15 27.92 75.25 Vmech. % 64.7 73.5 65.2 47.2 69.8 I.M.E.P. 91.4 119.4 92.3 59.2 107.8 B.S.F.C. lbs..487.432.488.7124.4521 B.H.P. hr. I.S.F.C. -lbA..315.3178.318.356.3152 I.H.P. hr..B.Th. % 26.8 30.2 27.23 18.2 28.9 I.Th. % 41.5 41.1 41.87 58.6 41.4 Qcooling B.T.U. 25420 32600 25850 22000 28520

TABLE XII. Engine Test Results Runs at Pm 36" Hg, Tm~ 80~F, Variable Speeds Run Number 98 99 101 105 16 10 1113 114 115 120 Pb ins. Hg. 29.25 29.25 29.25 29.25 29.25 29.58 29.53 29.53 29.53 29.16 Pm ins. Hg. 35.5 35.25 34.97 35.25 35.33 35.61 35.66 35.76 35.38 35.04 Tm ~F 78 82.5 86 88.5 89.5 76 82 83 86 79.8 N R.P.M. 993 1240 1555 1421 1001 1315 1257 1362 1769 1270 To ~F 155 172.5 186 185.5 176 172 173 178 193 177 Tw, exh. OF 170 181 182 191 204 169 186 183 190 182 TI1 ~F 188 201 205.5 215 231 181.5 209 208 214 207 TN2 ~F 178 189.5 190 193 205 178 192 190 196 189 TB2 ~F 174 182 182 183 191.5 174.1 185 183 186 182 T3 T ~F 132 145 153 154 148 144 146 148 158 146 ATc=(T3 -T 1) 6.07 6.05 6.15 7.64 10.75 3.99 6.8 6.96 6.26 7 T ~F 154 160 157 - 162 154 - - 152 - T ~F 158 165.5 162 160.5 170 157 168 174 158 162 C3! Texh. F 427 534 674 850 947 405 676 685 832 717 W lbs 111 137 174 159.2 111 149.2 138.5 156 199 140.2 a hr. w lbs. 2940 3642 4542 4170 2968 3860 3690 3990 5220 3720 c hr. Wf l 1lbs. 2.41 3.39 4.98 6.05 5.365 2.461 4.54 4.79 6.28 4.71 fuel hr. FRatio.0217.0246.0287.0379.0483.0165.0328.0307.0316.0336 A B.H.P. 4.55 6.61 9.41 13.43 12.39 3.13 9.5 10.17 11.4 10.35 B.M.E.P. 43.55 50.9 57.5 89.7 117.4 22.6 75.7 71 61.3 77.5 B.M.E.P. Corr. 42.05 49.5 56.2 88.7 116.7 21 74.55 69.8 60.1 76.4 Imech. % 59.7 60.8 59.85 71.7 79.4 39.9 68.4 67 58.2 69.9 I.M.E.P. 70.5 81.5 93.9 123.8 147 52.7 109 104.2 103.3 109.2 B.S.F.C. B.H1P.hr.548.528.542.4562.436.848.485.479.562.461 I.S.F.C. lbs..3265.3205.324.3262.346.338.3315.321.3265.322 I.H.P. hr. qB.Th. % 23.78 24.8 24.05 28.62 30 15.44 26.9 27.2 23.25 28.3 nI.Th. % 39.8 40.8 40.25 40 37.8 38.72 39.39 40.63 40 40.55 Qcooling B.h.U. 17850 22000 27870 31810 31900 15400 25100 27790 32600 26020

-127TABLE XIII. Engine Test Results Extra Runs Referred To In The Calculations Run Number 10 53 57 102 Pb ins. Hg. 29.34 29.52 29.68 29.25 Pm ins. Hg. 27.34 36.12 34.38 35.4 Tm ~F 72 81 77 88 N R.P.M. 1740 566 656 1760 TO ~F 184 153.5 148 194 Tw exh. ~F 194 184 168 184 Ti! OF 201 199.5 178 206 TY2 ~F 183 189 173 189.5 Tz3 ~F 175 182 170 181 3 Ts ~F 119.5 117.5 126 159 AT 3 = (T 1 -T ) 5.76 8.86 4.62 6.115 Tc2 ~F - 158 157 T. ~F 162 165.9 159.6 158.1 cs Texh. OF 842 509 318 746 Wa lh 137.1 63.9 71.5 197.5 *a hr. WC lbs, 5140 1730 1980 5232 hr. Wfuel hr. 5.46 2.16 1.217 5.72 E Ratio.0398.0338.01701.0289 A B.H.P. 6.46 4.2 1.615 9.71 B.M.E.P. 35.3 70.5 23.6 52.5 B.M.E.P. Corr. -69.4 22 51.2 rmech. % 46.4 73.5 46.5 54.5 I.M.E.P. 76.2 94.4 47.4 94 B.S.F.C. B.H.P...845.522.802.604 I. S.F.(C,".-H.Ph..392.384. 372295 B.Th. % 15.45 25 16. 21.6 rlI.Th. % 3355.5 34 55 59.7 Qcooling B.T.U. 29600 15540 9150 52000 hr.

TABLE XIV. Calculation of the Constant "C" in the Equation [2.11]. Run Number 53 57 62 74 89 95 96 97 Average water temperature ~F. 161.47 157.3 163.27 158.4 160.05 168 168 165 Average wall temperature ~F. 185.2 171.5 191 187.8 184.2 199 200 198.5 (Twall aver.-Twater aver.)OF 23.73 14.2 27.77 29.39 24.15 31 32 33.5 B.T.U. Qcooling B.T. 15340 9150 20670 22800 18350 24220 25200 24200 hr. a sq. B.T.U.o 445 444 514 534 524 540 544 497 r c sq. ft. hr.~F i at Twater average 0.958 0.99 0.945 0.975 0.965 0.905 0.905 0.93 Wc 1807 2000 2586 2550 2410 2705 2770 2617 (X)0.6 90 95.8 111 110 107 115 116 113 _A at average temperature 2.535 2.62 2.5 2.57 2.55 2.4 2.4 2.46 k (_)~0.4 1.45 1.47 1.443 1.459 1.454 1.42 1.42 1.434 k (W_)0.6 (c.)0.4 130.5 142.2 160.3 160.5 155.8 163.2 164.7 162 Ip. k

TABLE XV. Calculation of the Constant "C" in the Equation [2.11]. Run Number 10 98 101 102' 131 153 Average water temp. ~F 159.1 155 158.9 155.1 165.5 155.5 Average wall temp. ~F 185 179.2 186.5 186.5 198.5 187.5 (Twall aver.- Twater aver.)OF 25.9 24.2 27.6 51.4 55 32 Qocooling B.T.U./hr. 29600 17850 27870 52000 52200 28520 C B.T.U./sq.ft.OF.hr. 788 510 698 702 633 614 iL at Twater average 0.97 1.01 0.98 1.01 0.95 1.0 WC 5500 2910 4655 5170 5740 5580 ()0.6 171.5 120 159 170 139 155 C at average temp. 2.58 2.67 2.59 2.68 2.51 2.65 k (Ci)0.4 1.461 1.481 1.464 1.484 1.445 1.475 k (WC)0.6. (cC)0.4 250.5 177.8 255 252 201 199 [i- k

-130CALCULATION OF"C" IN EQUATION [2.11] 800'r 800 /6 6 400 700 L / 0 600 200 0 0 100 0 20 40 60 80 100 120 140 160 180 200 220 240 260 FIG. 57 \()^ (/ )

APPENDIX E CALCULATION OF THE MEAN WALL-TEMPERATURES FOR RUN NO. 95 1. Wall temperature over the whole cycle. 2. Wall temperature at any crank angle. 3. Mean wall temperature during the intake stroke. 1. Wall Temperature Over The W hole Cycle: The areas enclosing the gas consist mainly of the combustion chamber and the cylinder liner areas. a. The combustion chamber area equals 48.4 sq. ins. and its mean temperature was calculated from figure (22) = 243.6 F b. The cylinder liner area equals.-4.2 sq. ins. and its temperature at T.D.C. equals that of the combustion chamber and changes to lower values toward the B.D.C. This change was calculated from the readings of the three thermocouples measuring the liner outside surface temperature, and the temperature drop AT through the liner walls. D2 Qln rAT= 2jT kwL where Q = heat gained by water in the engine barrel = Wc x sp. ht. x (Tc-Tc1) = 2435 x 4 = 9740 B.T.U. hr. D2 = 5.063 ins. D1 = 4.5 ins. -131

-132kw: thermal conductivity of the wall = 27 L = stroke = 5.25 ins. 97401n 5.063 x 12.*A T = 4.5 = 15.6 ~F 2it x 27 x 5.25 This temperature drop was added to the reading of the thermocouple midway between the T.D.C. and the B.D.C., and the curve a b c d, figure (58), representing the liner inside wall temperature was drawn. 2. Wall Temperature At Any Crank Angle: The curve a b c d, figure (58) representing the liner temperature was extended to point e which represents the liner temperature at a point beyond the first piston-ring, and this temperature was assumed constant. At any crank angle, the temperature of the wall could be found by drawing a curve between point e and the temperature of the combustion chamber at the same angle. 3. Mean Wall Temperature During The Intake Stroke: The wall temperature during the intake stroke was found by using the method under item No. 2. The mean combustion chamber temperature during the suction stroke = 238~F. The mean temperature of the combustion chamber and liner, (figure 58) = 224~F. The mean area of the combustion chamber and liner, (figure 54) = 122.6 sq. in. During the intake stroke the air is in contact with the intake manifold, whose area equals 34.3 sq. ins., and at a temperature of 170~F, assumed equal to that of the cooling water in the cylinder head..'. Mean temperature of the walls during the intake stroke = 224 x 122.6 + 34.3 x 170 O 122.6 +212 OF 122.6 + 34.3

260 CALCULATION OF CYLIEDER WALLS U.n_~~~~~~~~~~ M1,~~~~EAN1 TEMIPERAIThE 0 FOR RUN NO. 95 * 250 o a Mean comb. chamber Temp. (All cycle) b 0. 240 Mean comb. chamber Temp.(intake strok 230 rS~~~~~~~~~~~~~~~~~~~~~~~~N 220 - 210 - t 200 d H ^ C i 190 - o i so C I~~ ~ ~ ~~~~~~~~~~~~~ >,c P7 ISBO - r?! d rd c)Cd 170 o H-i H-E * r-H p. ~r Ei P r 16 0 ~ ~ 5: ^ ~~~~160 g ~ ~~~~0 0 () C 0 0 ) A ~~~~~IP~L 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 Combustion chamber area JCylinder liner area Area sq in 48.4 74.2 FIG. 58

-134TABLE XVI CaM = Mean Coefficients of Heat Transfer and T E= Mean Effective Temperatures Run Number 1 94 30 95 96 62 97 70 lbs. Pm sq. in 14.08 14.48 15.82 16.02 17.59 19.30 18.86 20.56 Tm ~R 543 550 533.5 550 548 539.5 543 548.5 N R.P.M. 807 826 793 821 840 817 814 815 Ft. S - ec. 77 12.03 11.55 11.97 12.23 11.9 11.87 11.88 Intake 11.42 11.8 12.3 12.02 13.13 13.58 12.72 14.28 Compression 26.2 28 28.15 26.55 33.3 31.6 32.6 32.8 Expansion 51.8 68.1 52.7 70.3 76.4 69.4 77.4 79.6 Exhaust 16.1 18.6 14.9 18.56 18.95 16.9 17.37 17.85 Total Cycle 26.38 31.63 27.01 31.86 35.45 32.87 35.02 36.13 Intake.688.74.767.796.811.821 ~797.895 o Compression 2.3 2.85 2.o7 2.415 3.4 3.14 3.38 3.23 _,Expansion 8.97 15.4 9~74 15.5 18.18 14.88 17.86 17.78 F Exhaust 1.535 2.415 1.32 2.48 2.48 1.641 2.12 2.27 b Total Cycle 3.373 5.351 3.624 5.298 6.218 5.121 6.04 6.044 M.E. R 1278 1690 1342 1660 1755 1557 1722 1670 (M Calculated 25.85 30.2 27.22 32.7 34.8 32.5 34.30 36.95 Calculated 380 1718 1381 1750 1738 1504 1646 1700 I.M.E.P 61.2 112.5 70.0 132 142.2 107.8 132.5 152 lbs. sq. in.

-135TABLE X7II = Mean Coefficients of Heat Transfer and TTl = TMean Effective Temperatures Run Number 84 89 72 74 127 131 124 133 lbs. Pm s. in. 20.24 20.5 21.68 21.6 17.58 17.42 17.4 17.46 T ~R 546 545 543.5 548 605 601 657 664 m N R.P.M. 805 775 823 832 1216 1204 1231 1214 S Ft. |S Ft. 11.72 11.3 12 12.12 17.72 17.58 17.97 17.71 sec. Intake 14.18 13.97 14.58 13.97 15.33 15.22 15.9 15.7 Compression 36.6 34.3 35.28 33.5 36 34.9 34 34.8 Expansion 54.6 60 57 68.4 83.1 87.7 81.7 78.4 |Exhaust 14 14.42 14 16 21.35 21.57 22.18 20.55 Total Cycle 29.85 30.67 30.215 32.97 38.95 39.85 38.45 37.33 Intake.902.87.893.912.975 1.0 1.085 1.083. /ompression 3.5 3.31 3.48 3.25 3.882 3.725 3.615 3.79.xpansion 9.0 11.35 8.94 13.43 19.4 22.0 19.6 18.15 4 xhaust 1.082 1.30 1.068 1.503 2.88 2.96 3.13 2.64 I Total Cycle 3.621 4.21 3.595 4.774 6.784 7.421 6.856 6.416 TME. OR 1212 1372 1190 1448 1740 1862 1782 1720 Calculatd 29 30.75 29.82 33.32 38.60 39.20 39.33 38.57 CalLculated TI.E. oR 1202 1352 1178 1422 1757 1792 1852 1790 Calculated I.M.EP*. 57.3 83 58 98.6 120.8 129.8 119.4 107.8 lbs. sq. in.

-136TABLE XVIII COMBUSTION CAIAMBER WALL TEMPERATURES CALCULATED A;ID MEASURED Pm F I.M.E.P. Tw. Tg Run inches g Tm N A lbs. per calcu- measured Error % Error No. absoldte OF R.P.M. Ratio sq. in. lated OF OF OF 2 1 28.75 83 807.0261 61.2 192.5 212.8 20.3 9.5 a 2 28.78 84 800.0463 101 225.2 231.9 6.7 2.9 194 29.55 90 826.0515 112.5 273.8 245 -28.8 -11.7 4 28.56 72 810.0528 116.2 251 258.3 7.3 2.8 81 32.84 86 796.0205 56.4 198.5 219 20.5 9.4 -g 93 32.55 87.5 832.0455 118.2 261.8 237.9 -23.9 -10 O M 136 32.3 87 790.0444 121.5 237.2 231.6 - 5.6 -2.4 | 595 32.62 90 821.0511 132 249.3 243.6 5.7 2.34 r 37 31.93 77 830.0546 141.6 242.7 241.6 -loi -0.5 8 47 35.12 73 828.0179 51.3 188.2 200.4 12.2 6.1 00 O co aO 51 35.25 80.5 850.0431 129 240.8 227.2 -13.6 -6 152 35.17 81 847.0519 145.7 250.8 229.7 -21.1 -9.2 1 83 38.54 87 820.0174 53.3 200.4 208.5 8.1 3.9 4| 97 38.47 83 814.0412 132.5 250.1 237.9 -12.2 -5.1 64 39.24 82 807.0424 146.5 241.8 246.2 4.4 1.8 71 39.27 90 789.0448 149 255.2 260.7 5.5 2.1 66 41.47 76 827.0154 54.8 193.1 206.7 13.6 6.6 l 90 41.8 72.5 837.0336 116 234.3 245 10.7 4.4 "69 41.72 86.5 805.035 127 240.5 252.2 11.7 4.6 70 41.97 88.5 815.0434 152 262.5 260.4 -2.1 -0.8 72 44.25 83.5 823.0154 58 186.6 211.1 24.5 11.6 L 7\ I 74 44.05 88 832.0257 98.6 220.2 244.9 24.7 10.1 175 44.55 89.5 815.030 118.6 234.2 254.6 20.4 8 I 76 44.70 90 815.0357 135.2 237.7 250.4 12.7 5

-137TABLE XIX COMBUSTION CHAMBER WALL TEIIPERATURES CALCULATED AND MEASURED Pm F I.M.E.P. T T Run inches Hg Tm N A Ibs. per calu- measured Error o Error C No. absolute OF R.P.M Ratio sq. in. lated OF OF OF P 98 35.5 78 993.0217 70.5 203.2 232.8 29.6 12.7 106 35.33 89.5 1001.0483 147 255 259.6 4.6 1.8 cd 99 35.25 82.5 1240.0246 81.5 215.7 235.4 19.7 8.4 113 35.66 82 1257.0328 109 229.8 236.9 7.1 3 0 0 Co. 120 35.04 79.8 1270.0336 109.2 225.2 25.5 31.3 12.2 H5 110 35.61 76 1315.0165 52.7 190.9 187.9 -3 -1.6 114 35.76 83 1362.0307 104.2 232.8 245 12.2 5 0 105 35.25 88.5 1421.0379 123.8 233 257.2 24.2 9.4 F 1l1 34.97 86 1555.0287 93.9 220.2 243 22.8 9.4 115 35.38 86 1769.0316 103.3 218.7 248 29.8 11.8 o 104 35.45 90.5 1197.0362 122 243.c 260.5 16.7 6.4 Co n136 35.48 82.5 1223 -o38 123.1 226.6 264 27.4 10.4 EH 130 35.83 80 1190.0451 145.9 240.8 246.2 5.4 2.2 ~ 128 35.78 144 1204.0239 70 215.4 223 7.6 3.4 "o126 35.72 141 1216.0311 96.7 230.2 240.9 10.7 4.4. — r 1127 35.35 145 1216.0397 120.8 239.4 242.7 3.3 1.4 H 131 35.56 141 1204.0468 129.8 251.8 247.5 -4.3 -1.7 123 35.64 193 1212.0334 91.4 238.4 241.5 3.1 1.3 q 1125 35.71 198 1212.0338 92.3 238.8 248 9.2 3.7 133 35.62 204 1214.0367 107.8 242.7 228.5 -14.2 -6.2

-138TABLE XX CALCULATED THERMAL LOADINGS AND MEASURED HEAT LOSSES TO COOLING WATER U Q Q Run B.T.U. Calculated Measured Error % Error No. hr. sq.ft. ~F B.T.U. B.T.U. B.T.U. hr. hr. hr. 1 24.36 13350 13900 350 2.52 2 26.4 19100 18900 -200 -1.06 9' 94 28.1 21270 21650 380 1.75 p- 4 27.6 21250 22400 1150 5.15 81 24.65 11720 15000 3280 21.85 93 29.4 21900 21650 -250 -1.15 2S 36 29 22200 21300 -900 -4.23 E-I 1 95 30.4 23900 24220 320 1.32 2 37 30.4 25050 26280 1230 4.68 o 47 25.3 10670 12420 1750 14.1. oM 51 31.4 22700 24600 1900 7.73 0~ II 8 (~ 52 31.95 25600 27800 2200 7.9 CId 83 26.3 10770 12300 1530 12.4 o N 97 31.63 22860 24200 1340 5.55 4) 2 "1 64 32.7 24600 26500 1900 7.16 o pM 71 32.7 25370 28800 3430 11.9 66 26.9 10470 12800 2330 18.2 (M 90 31.38 19650 21200 1550 7.3 69 31.95 21900 24000 2100 8.75 70 33.9 25750 29550 3800 12.87 72 27.8 11200 12720 1520 11.94 u 74 30.93 17520 22800 5280 23.2 " 75 32.4 20400 22400 2000 8.95 m 76 33.5 22900 25500 2600 10.2

-159TABLE XXI CALCULATED THERMAL LOADINGS AND MEASURED HEAT LOSSES TO COOLING WATER U Q Q Run B.T.U. Calculated Measured Error % Error No. hr. sq.ft. OF B.T.U. B.T.U. B.T.U. hr. hr. hr. d 98 28.5 14680 17850 3170 17.75, 106 34.1 27800 31900 4100 12.85 H 99 31.6 17800 22000 4200 19.1 X 113 33.9 23100 25100 2000 7.97 & 120 33.86 22830 26020 3190 12.25 110 29.95 12670 15400 2730 17.72 114 34.4 22200 27790 5590 20.1 { 105 36.45 26900 31810 4910 15.45 M 101 35.1 21880 27870 5990 21.45 p 115 37.45 24770 32600 7830 24 o 104 34.37 24850 24750 100 0.4 co 11136 34.56 25300 27620 2320 8.4 H 130 36 28750 32250 3500 10.85 0 i 128 31.6 18250 22550 4300 19.1 126 33.7 23500 25300 1800 7.1 H 127 35.62 28500 30580 2080 6.8 " ~131 36.4 29650 32200 2550 7.92 11 0 123 34.15 25000 25420 420 1.65 125 34.23 25500 25850 350 1.35 H 133 35.6 29300 28520 -780 -2.74

APPENDIX F SCAVENGING AIR FLOW-RATE For the engine used in experimental work, the opening of the inlet valve and the closing of the exhaust valve allowed a period of overlap of 36 crank angles. Air flowing during this period through the exhaust valve does not take part in the combustion process. For the purpose of comparison between the operating conditions studied it was assumed that the air flow rate is function only of the square root of the difference in pressure head between the inlet and exhaust manifolds. While the engine was at rest the air pressure in the intake surge tank was kept constant at 6" Hg above the atmospheric, while the back pressure was atmospheric, the air flow rate was measured for different crank angles over the scavenging period. The results are shown in the following table and in figure (59). Crank angles 6.5 5.5 4 3.7 2.7 0.9 before T.D.C. Flow rate lbs..00495.0071.01005.0112.0127.0185 sec. Crank angles 0.6 2.4 3 6.8 8.2 9.7 11.3 after T.D.C. Flow rate lbs..0224.0256.0258.0196.0049.00115.00075 sec. From figure (59) the average flow rate of the scavenging air, -140

-141(at A P = 6" Hg. and Tm = 77~F) = 0.764 x 10-2 lbs. per sec. For any run with the air at Pm and Tm the average flow rate during the scavenging period was calculated from W = 0.764 x 10-2 (m-i ) p se537. W = 0.764 x 102 ix lbs. per sec. a6Tm

c'J 0 x o SCAVENGING AIR FLOW -RATE DIAGRAM U) i~~~s~ ~FOR n-I gn~~~ ~ Ta = 770F IAP=6 HgE Average Flow Rate = 0,762 x 102 lb s. B -3.0 sec. C2.5/V 2.0 1.5 t.0 -0.5 16 14 12 10 8 6 4 2 0 2 4 6 8 10 12 14 16 18 20 CRANK ANGLES BEFORE T.D.C. T.D.C. CRANK ANGLES AFTER T.D.C. FIGURE 59

APPENDIX G SAMPLE CALCULATIONS FOR THE EFFECT OF AFTERCOOLING Assumed Operating Conditions: Aftercooler effectiveness, e = 50% Air pressure at the intake manifold = 45 ins. Hg. Pressure drop between compressor and engine = 0.5. sq.in. P1: Pressure at compressor inlet = 14.7 lb. sq. in. T1: Temperature at compressor inlet = 540~R c: Compressor efficiency = 80% Calculations: h*1 B.T.U. lb. Pr 1.386 22.5 P 1.386 x 2 2.12 r2 14.7 h, 145.8 B.T.U. Ah isentropic = h2t- h = 145.8 - 129.1 = 6.7 B.T.U.,h h lb. h2-hl 2 1 = 20.86 rlc h2 149.96 B.T.U. 1lb. T2 = 627 % T2-Tm c: Aftercooling effectiveness = T-T 50 % 2 1'. Tm = 583.5 ~R Tm Correction factor for the I.M.E.P., T = 0.925 TKeenan and Kaye Tables, Reference No. 20. -143

-144The power output of the turbocharged engine was calculated from the I.M.E.P. values of 45 ins. Hg. of figure (25), and the correction factor 0.925. The calculations will be continued for an output of 140 lbs. per sq. in. TM.E.calculated from equation [7.1] = 1682 ~R a calculated from equation [7.2] = 37.3 B.T.U. M hr. sq. ft. ~F kw = 27 B.T.U. hr. sq. ft.(^) Equating in equation [2.21] we get Tp.g.: Piston-Top maximum temperature = 965 ~F For the calculations of the intensity of thermal loads the following conditions were considered: TC: cooling water temperature = 160 ~F W^: cooling water flow rate = 2400 lbs B.T.Iu. Substituting in equations [2.11], we get ac = 500 hr. sq. ft. F am, from equation [7.5] = 0.877 ac, from equation [7.6] = 0.782 x, from equation [7.7] =.282 in. By substitution in equation [2.15], the intensity of thermalload = 36490 q.ft. hr.sq.ft.

TABLE XXII PISTON-CROWN TEMPERATURES, AND INTENSITY OF THERMAL LOADS E = 0% PM TM Piston-Crown Maximum Temperatures Intensity of Thermal Loads ins. Hg. 0R OF B.T.U. sq.ft. hr. I.M.E.P. lbs. 60 80 100 120 140 150 60 80 100 120 140 150 30 540 662 769 857 958 1040 1081 17985 22500 26890 31300 35720 37950 33 565 669 777 873 965 1051 1092 18770 23570 28050 32790 37350 39650 36 581 666 775 870 961 1046 1085 19100 23970 28650 33330 38050 4o46o 39 597.5 664 771 867 958 1045 1085 19480 24420 29240 34050 39010 41390 42 613.5 661 768 865 995 1042 1083 19770 24870 29780 34670 39690 42160 45 627 658 764 861 953 1038 1076 20060 25210 30230 35230 40350 42830

TABLE XXIII PISTON-CROWN TEMPERATURES, AND INTENSITY OF THERMAL LOADS E = 50% Pm Tm Piston-Crown Maximum Temperatures Intensity of Thermal Loads ins. Hr. OR ~F B.T.U. _______________________________sq.ft. hr. I.M.E.P. lbs. sq. in. 60 80 100 120 140 150 60 80 100 120 140 150 30 540 662 769 864 954 1040 1081 17985 22500 26890 31300 35720 37950 33 552.5 654 759 854 945 1029 1069 18090 22700 27140 31700 36260 38440 36 560.6 642 746 840 929 1012 1052 18050 22700 27250 31720 36390 38580 39 568.8 631 733 825 913 995 1035 18000 22640 27220 31780 36420 38670 42 576.8 619 722 813 900 982 1020 17940 22660 27270 31890 36490 38770 45 583.5 608 709 799 885 965 1003 17820 22550 27140 31810 36490 38940

TABLE XXIV PISTON-CROWN TEMPERATURES, AND INTENSITY OF THERMAL LOADS c = 100% P iTM Piston-Crown Maximum Temperatures Intensity of Thermal Loads ins. Hg. OR OF B.T.U. sq. ft. hr. oM.E.P. lbs. sq.in. 60 80 100 120 140 150 60 80 100 120 140 150 | 30 540 662 769 864 954 1040 1081 17985 22500 26890 31300 35720 37950 33 540 638 742 835 924 1006 1046 17470 21995 26320 30760 35080 37330 36 540 617 718 809 895 976 1014 16990 21480 25820 30180 34500 36740 39 540 597 696 784 869 948 986 16530 20980 25250 29600 33850 36080 42 540 578 675 762 844 921 959 16050 20490 24740 29030 33380 35480 45 540 563 657 743 823 899 937 15710 20090 24360 28530 32880 35020

Pm= 30" 1100 33" 36" 42"1 45" Tp.g.: PISTON-CROWN TEMPERATURE VS. I.M.E.P. 3 1000 FOR DIFFERENT MANIFOLD PRESSURES =00 100 900 800 O 0 00 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 180 I.M.E.P. sq.in. FIGURE 600 /

P:30" 39000 33 36 39 +742 35000 q: INTENSITY OF THERMAL LOADS VS. I.M.E.P. 45 FOR DIFFERENT MANIFOLD PRESSURES 7 31000 0 27000 1 23000 19000 15000 — IIII 0 20 40 60 80 100 120 140 160 I.M.E.P. lbs./sq.in. FIGURE 6i

30" Hg 33 1100 I 36 Tp.g. PISTON-CROWN TEMPERATURE VS. I.M.E.P./ 2 FOR DIFFERENT MANIFOLD PRESSURES 45 50%500L l —--- l —— l —-— l —— l —--- l ---- l l —45 =FIGURE 62 1000 B 800 kl og0 700 600 500 0 20 40 60 80 100 120 140 160 I.M.E.P. lbs./sq.in. FIGURE 62

45" 39000 30 q: INTENSITY OF THERMAL LOADS VERSUS I.M.E.P. FOR DIFFERENT MANIFOLD PRESSURES 35000 +; ~ = 50% 31000 27000 H 0 g 23000 19000 15000 -LI I I I 0 20 40 60 80 100 120 140 160 I.MoE.P. lbs./sq.in. FIGURE 63

1100- 33" Hg — t 45': Hg Tp.g.: PISTON-CROWN TEMPERATURE VS. I.M.E.P. 1000 FOR DIFFERENT MANIFOLD PRESSURES e =0 900 800 80 0 CO) 700 600 500 I I I I I I I I I I 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 I.M.E.P. lbs./sq.in. FIGURE 64

45" Hg 43000 42" IIg / 9" Hg 41000, 36X/ 56" Hg + 3 55? Hg 39000 q: INTENSITY OF THERMAL LOADS VS. I.M.E.P. // / 50" Hg FOR DIFFERENT MANIFOLD PRESSURES 37000 35000 33000 rq 31000 / / 29000 27000 I o 25000 23000 21000 19000 17000 15000 I I I I I I I I I I I I 0 10 20 30 40 50 60 70 80 90 100 110 120 130 140 150 160 170 I.M.E.P. lbs./sq.in. FIGURE 65

43000 39000 q: INTENSITY OF THERMAL LOADS VS. I.M.E.P. FOR DIFFERENT EFFECTIVENESS.35000 Pm = 45" g CH r 31000 0 27000 0 H 2000 23000 / 19000 15000 I I 0 20 40 60 80 100 120 140 160 FIGURE 66 I.M.E.P. lbs./sq.in.

APPENDIX H CORRECTION FACTOR FOR THE. B.M.E.P. The effect of the back pressure on the B.M.E.P. was not fully investigated. Some tests were made at different bick pressures, and the B.M.E.P. was not affected until the back pressure was boosted to valves near the inlet manifold pressure. This indicated that the friction load, on the single cylinder engine used for the tests, was not very sensitive to the increase in the back pressure and the drop in the power output was rather due to the poor scavenging efficiency at these conditions. (5) P-Pis uo te r o exh. The scavenging efficiency is fuc-tion on the ratio mPh To get the drop in the B.M.E.P. due to increase in the back pressure for an equal scavenging efficiency, a value of, C.04, was used for the above ratio. This low valve was chosen, because of the relatively low supercharging pressures reached in the tests. Few runs were made at an average intake manifold pressure of 35.6" Hg. The back pressure was changed from atmospheric to 54.2, with the F constant, (Table 9). The drop in B.M.E.P. is shown in figure (67). -155

-156CORRECTION FACTOR FOR THE B.M.E.P. *H 4 C. o01.02.03.04.05.06 F Ratio A FIGaURE 67

APPENDIX I ENGINE SPECIFICATIONS Manufacturer: Nordberg Mfg. Co., Milwaukee, Wisconsin, U.S.A. Number of cylinders: 1 Cycle: 4 strokes per cycle Compression ratio: 14.5 Bore: 4 1/2 inches Stroke: 5 1/4 inches Piston Displacement: 35.48 cubic inches. Combustion Chamber Type: Energy Cell Engine Timing, Crank Degrees Fuel pump port closing: 30~ before T.D.C. Injection begins (approximately): 19~ before T.D.C. Exhaust valve opens: 45~ before B.D.C. Exhaust valve closes: 20~ after T.D.C. Intake valve opens: 16~ before T.D.C. Intake valve closes: 38~ after B.D.C. -157

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UNIVERSITY OF MICHIGAN 3 9015 03026 6384i 0302 6384n EQ