”‘71 (.10 ,1, ,I #1! o. .L'AI If A STUDY OF THE INFLUENCE OF DIFFERENTIAL PRESSURES IN INTAKE-MANIFOLD BRANCHES UPON DEVELOPED ENGINE HORSEPOWER By William Ea rl Bishop AN A BSTRACT Submitted to the School of Graduate Studies of Michigan State College of Agriculture and Applied Science in partial fulfillment of the requirements for the degree of MASTER O F SCIENCE Department of Mechanical Engineering Year 1955 d J . ~-- 1’1. ,, , 7” Approved V A?! ace...“ (/4 \ ( /.. A A5 THESIS ti...‘ AIL-bl , I ‘94. if WILLIAM E. BISHOP ABSTRACT It is difficult to find. any major engine component or assembly which can be as indeterminate in design as the induction system. Although practices in induction system design have changed very little, much new information is still necessary. The main object in this experiment is a study of the relation between intake-manifold branch pressure differential and developed engine horsepower. Differential pressures are measured for each cylinder between the base of the carburetor and the intake port. Indicated horsepower is recorded for the same Operating conditions. Four different manifold systems were tested with the throttles at fixed positions and increments in r.p.m. were varied by a change in load. The performance characteristics of the engine for each intake manifold system may help to distinguish the most effective system with particular emphasis on output. There is not a specific variation or a definite relationship between pressure differential and deve10ped engine horsepower. The re are other variables present that affect the state of distribu- tion. Load, r.p.m. , and flow rate are the most important of these. The deveIOped horsepower will be more uniform when the load and r.p.m. are high. At low r.p.m. and high load, the WILLIAM E. BISHOP ABSTRACT developed horsepower will be higher per cylinder, if the pressure differential is high with the low pressure end of the differential located at the intake port. This does not imply that the developed horsepower will only be a maximum under those conditions. The same high horsepower can be obtained with a low pressure differ- ential with the valve having the higher pressure, This is most likely due to the liquid end of the fuel particles that enter the cylinder at that time. The liquid fuel particles do not have a definite flow pattern. Under assumed low volumetric efficiency conditions, the minimum velocity in the largest intake manifold branch is slightly less than 10 feet per second. The minimum carrying velocity for a liquid fuel of the Specific gravity of gasoline is 2 feet per second. At high flow rates, the pressure differential has little effect on the develOped horsepower, and the ramming effect is alm05t always present. When the ramming effect is. not present, the absolute pressure is high enough to obtain an adequate flow to the cylinder. The load factor shifts the point where ram initiates. As the load is increased the r.p.m. where ram starts is reduced. WILLIAM E. BISHOP ABSTRACT The performance curves for each induction system vary little for engine Speeds below 3800 r.p.m. Further increases in r.p.m. exhibit a little more variation. In brief summary, when ramming effect is present, the de- ve10ped horsepower will be high per cylinder. When ramming ef- fect is absent, the horsepower is usually lower. Under an absolute pressure analysis, if the absolute pressure is high at the intake port, the deveIOped horsepower will be high. A STUDY OF THE INFLUENCE OF DIFFERENTIAL PRESSURES IN INTAKE-MANIFOLD BRANCHES UPON DEVELOPED ENGINE HORSEPOWER By WILLIAM EARL BISHOP A THESIS Submitted to the School of Graduate Studies of Michigan State College of Agriculture and Applied Science in partial fulfillment of the requirements for the degree of MASTER OF SCIENCE Department of Mechanical Engineering 1955 ACKNOWLEDGMENTS The author wishes to express his sincere appreciation to Dr. L. L. Otto for his invaluable contributions to this investiga— tion, and to the engineering staff for their assistance in many phases of the experimental work. ii 354801 F“ Int-.T— .‘P' AUTOBIOGRAPHY The author graduated from Romulus High School in 1944 and enrolled in Michigan State College for the summer session in 1944. Eighteen months were then Spent in the Navy as a M.M.R. (Machinist Mate Refrigeration). Re—entering Michigan State College in the winter term of 1947, he attended college intermittently until he received his "B.S. " degree in the spring of 1952, in the power option of Mechanical Engineering. He. entered the Graduate School in the winter term, 1953. This thesis is in partial fulfillment of the requirements for the "M.S." degree, to be granted in June, 1955. iii TABLE OF CONTENTS IIIIIIIIIIIIIIIIIIIIIIIIIII CHOICE OF OPERATING CONDITIONS ............... P R0 CE DURE ...... DESCRIPTION OF APPARATUS .................... DESCRIPTION OF TEST RESULTS .................. Discussion of Figures 1~22 ..................... Discussion of Figures 23—56 .................... Discussion of Figures 57-100 ................... APPENDIXES . . . . Appendix A-l, Performance Data . . . . . . . . ......... Appendix A-Z, Data for I.HP./Cyl. and Friction ...... Appendix B-‘l, The Determination of Engine Characteristics . . ooooooooooooooooooooooooooo Appendix B-Z, The Determination of I.HP. /Cyl. and Total Friction 13 18 21 26 Z7 Z7 33 39 152 153 163 171 181 Appendix Assumed Appendix Appendix C, Engine Dimensions and Tolerance Constants D, Selected E, Pictures References ................. ooooooooooooooooooooooooo Page 189 192 194 I. tie Figure H 10. 11. 12. 13. 14. 15. LIST OF FIGURES Perfgrfmavnfiget Curve 5- 1B. HP. , I. HP); F. HPLJ‘Iorque 1/8 throttle Standard Manifold .............. 1/4 throttle — Standard Manifold .............. 1/2 throttle - Standard Manifold .............. 3/4 throttle — Standard Manifold .............. Full throttle - Standard Manifold ..... . . . . . . . . 1/16 throttle - Dual Carburetor Manifold . . . . . . . 1/8 throttle - Dual Carburetor Manifold ........ 1/4 throttle - Dual Carburetor Manifold ........ 1/2 throttle - Dual Carburetor Manifold Full throttle - Dual Carburetor Manifold ....... 1/8 throttle - Quad. Manifold--Secondary throttle Opens at 1/2 Primary ..................... 1/4 throttle - Quad. Manifold-«Secondary throttle Opens at 1/2 Primary ..................... 1/2 throttle - Quad. Manifold-~Secondary throttle Opens at 1/2 Primary . . .g .................. 3/4 throttle - Quad. Manifold--Secondary throttle Opens at 1/2 Primary ..................... Full throttle - Quad. Manifold-~Secondary throttle Opens at 1/2 Primary ..................... vi Page 49 50 51 52 53 54 55 56 57 58 59 60 61 62 63 Figure Page 16. 1/8 throttle - Quad. Carburetor Adapted to Standard Manifold ....................... ()4 17. 1/4 throttle - Quad. Carburetor Adapted to Standard Manifold ........................ 65 18. 1/2 throttle - Quad. Carburetor Adapted to Standard Manifold ....................... 66 19. Full throttle - Quad. Carburetor Adapted to Standard Manifold ........................ 67 20. 1/4 throttle - Quad. Manifold——Secondary throttle Opens at 1/4 Primary ..................... 68 21. 1/2 throttle - Quad. Manifold--Secondary throttle Opens at 1/4 Primary ..................... 69 22. Full throttle - Quad. Manifold--Secondary throttle Opens at 1/4 Primary ..................... 70 PerformanceACurvesfi-n-B. M. E. P..L M. E. , B..S. FLC. LMi./gal. 23. 1/8 throttle - Standard Manifold ....... . ..... 72 24. 1/4 throttle - Standard Manifold . . . . . . ....... 73 25. 1/2 throttle - Standard Manifold ............. 74 26. 3/4 throttle - Standard Manifold ............. 75 27. Full throttle - Standard Manifold ............. 76 28. 1/16 throttle - Dual Carburetor Manifold ....... 77 29. 1/ 8 throttle - Dual Carburetor Manifold ........ 78 30. 1/4 throttle - Dual Carburetor Manifold ........ 79 31. 1/2 throttle - Dual Carburetor Manifold ........ 80 vii Figure 32. 33. 34. 35. 36. 37. 38. 39. 40. 41. 42. 43. 44. 45. Full throttle - Dual Carburetor Manifold. 1/8 throttle - Quad. Manifold——Seeondary throttle Opens at 1/2 Primary .......... . .......... 1/4 throttle - Quad. Manifold-»Secondary throttle Opens at 1/2 Primary ..................... 1/2 throttle - Quad. Manifold--Seoondary throttle Opens at 1/2 Primary ................ . . . . . 3/4 throttle - Quad. Manifold--Secondary throttle Opens at 1/2 Primary ..................... Full throttle - Quad. Manifold--Secondary throttle opens at 1/2 Primary ..................... 1/8 throttle - Quad. Carburetor Adapted to Standard Manifold ........................ 1/4 throttle - Quad. Carburetor Adapted to Standard Manifold ........................ l/z throttle - Quad. Carburetor Adapted to Standa rd Manifold ........................ Full throttle - Quad. Carburetor Adapted to Standard Manifold ........................ Comparison of Curves Comparison of I.HP. Curves - Standard Manifold Comparison of B.HP. Curves - Standard Manifold . Comparison of Torque Curves - Standard Manifold . Comparison of I.HP. Curves - Dual Carburetor Manifold .............................. v iii Page 81 82 83 84 85 86 87 88 89 90 92 93 94 95 Figure 46. 47. 48. 49. 50. 51. 52. 53. 54. 55. 56. 57. 58. 59. Comparison of B.HP. Curves - Dual Carburetor Manifold Comparison of Torque Curves - Dual Carburetor Manifold ....... . Comparison of I.HP. Curves - Quad.Manifold-- Secondary throttle opens at 1/2 Primary Comparison of B.HP. Curves - Quad. Manifold-- Secondary throttle opens at 1/2 Primary ....... Comparison of Torque Curves - Quad. Manifold—- Secondary throttle Opens at 1/2 Primary ....... Comparison Of I.HP. Curves - Quad. Manifold-- Secondary throttle Opens at 1/4 Primary ....... Comparison of B.HP. Curves - Quad. Manifold-- Secondary throttle Opens at 1/4 Primary ..... . . Comparison of Torque Curves ~ Quad. Manifold-- Secondary throttle Opens at 1/4 Primary ....... Comparison Of I.HP. Curves - Quad. Carburetor Adapted to Standard Manifold ................ Comparison Of B.HP. Curves - Quad. Carburetor Adapted to Standard Manifold . . . . . . . ......... Comparison of Torque Curves - Quad. Carburetor Adapted to Standard Manifold ................ DflfergguaLPmsstg‘res and I.HP. I.HP. — 1/8 throttle ~ Standard Manifold . . . . . . . "H20 — 1/8 throttle - Standard Manifold ....... I.HP. - 1/4 throttle - Standard Manifold ....... ix Page 100 101 102 103 104 105 106 108 109 110 Figure 60. 61. 62. 63. 65. 66. 67. 68. 69. 70. 71. 72. 73. 74. 75. 76. 77. 78. 79. "H20 - 1/4 throttle - Standard Manifold ....... I.HP. - 1/2 throttle - Standard Manifold ....... "H20 - 1/2 throttle - Standard Manifold ....... I.HP. - 3/4 throttle - Standard Manifold ....... "H20 - 3/4 throttle - Standard Manifold ....... I.HP. - full throttle - Standard Manifold ....... "HZO - full throttle - Standard Manifold . . . . . . . I.HP. - 1/16 throttle - Dual Carburetor Manifold "HZO - 1/16 throttle - Dual Carburetor Manifold I.HP. - 1/8 throttle - Dual Carburetor Manifold . . "H20 - 1/8 throttle - Dual Carburetor Manifold I.HP. - 1/4 throttle - Dual Carburetor Manifold . . "H20 - 1/4 throttle - Dual Carburetor Manifold I.HP. - 1/2 throttle - Dual Carburetor Manifold . . "H20 - 1/2 throttle - Dual Carburetor Manifold I.HP. - full throttle - Dual Carburetor Manifold . "Hz - full throttle - Dual Carburetor Manifold . . I.HP. - 1/8 throttle - Quad. Manifold-~Secondary throttle Opens at 1/2 Primary . . . . . . ...... . . . "H O - 1/8 throttle - Quad. Manifold-~Secondary throttle Opens at 1/2 Primary ............... I.HP. " 1/4 throttle — Quad. Manifold-«Secondary throttle Opens at 1/2 Primary . . . . . . . . . ...... X Page 111 112 113 114 115 116 117 118 119 120 121 122 123 124 125 126 127 128 129 130 Figure Page 80. “H O - 1/4 throttle - Quad. Manifold--Secondary throttle Opens at 1/2 Primary ............... 131 81. I.HP. - 1/2 throttle - Quad. Manifold--Secondary throttle Opens at 1/2 Primary . . . . ........... 132 82. "H O - 1/2 throttle - Quad. Manifold--Secondary throttle Opens at 1/2 Primary ............... 133 83. I.HP. - 3/4 throttle - Quad. Manifold—-Secondary throttle opens at 1/2 Primary . . . . . .......... 134 84. "H O - 3/4 throttle - Quad. Manifold--Secondary throttle opens at 1/2 Primary ............... 135 85. I.HP. - full throttle - Quad. Manifold——Secondary throttle Opens at 1/2 Primary ...... . ........ 136 86. "H20 - full throttle - Quad. Manifold-~Secondary throttle Opens at 1/2 Primary ...... . . . . . . . . . 137 87. I.HP. - 1/8 throttle - Quad. Carburetor Adapted to Standard Manifold ...................... 138 88. "H o - 1/8 throttle — Quad. Carburetor Adapted to Standard Manifold ...................... 139 89. I.HP. - 1/4 throttle - Quad. Carburetor Adapted to Standard Manifold . . . . . . ........ . ....... 140 90. "H o - 1/4 throttle - Quad. Carburetor Adapted to tandard Manifold ...................... 141 91. I.HP. - 1/2 throttle - Quad. Carburetor Adapted to Standard Manifold . . . . . ..... . ........... 142 92. "H o - 1/2 throttle - Quad. Carburetor Adapted to Standard Manifold ........... . ....... . . . 143 93. I.HP. - full throttle - Quad. Carburetor Adapted to Standard Manifold ..... . . . .............. 144 xi Figure 94. 95. 96. 97. 98. 99. 100. "H O - full throttle -— Quad. Carburetor Adapted tO Standa rd Manifold ...................... I.HP. - 1/4 throttle - Quad. Manifold--Seoondary throttle Opens at 1/4 Primary . . . ....... . . . . . "H O - 1/4 throttle - Quad. Manifold--Secondary throttle opens at 1/4 Primary ............... I.HP. «- 1/2 throttle - Quad. Manifold—-Secondary throttle Opens at 1/4 Primary . . . ............ ”H o - 1/2 throttle - Quad. Manifold-"Secondary throttle Opens at 1/4 Primary ..... . . . . . . . I.HP. - full throttle - Quad. Manifold--Secondary throttle Opens at 1/4 Primary ............... "H O - full throttle - Quad. Manifold--Secondary throttle Opens at 1/4 Primary ...... . ........ Page 145 146 147 148 149 150 151 INTRODUCTION It is difficult to find any major engine component or assembly which can be as indeterminate in design as the induction system. Although practices in induction system design have changed very little, much new information is still necessary. The design engi- neer cannot say, ”These are the requirements we must meet, and the induction system will be designed in this way. " One cannot predetermine with any exactness the configuration or size of an induction system. This unit is usually built by trial and error. In design it is lmown that a carburetor directly connected to an intake port in a single cylinder engine will give higher horse- power per cubic inch than a multicylinder engine fitted in a like manner. This might be attributed to the elimination of an intake manifold, where an increase in charge resistance is encountered. This may not be the situation, but little material is presented on induction systems in the literature. The few articles that are pre— sented are old with reference to updraft carburetion, where riser velocities have a minimum for good fuel entrainment. Charge re- bound and accompanying precipitation Of fuel within the manifold are important factors in charge resistance. Practically all modern 1 engines use downdraft induction systems and it is the Opinion of this author that the requirements for updraft carburetion do not become requirements for downdraft carburetion. . |..! Q. HISTORY The function of a carburetor is to proportion the fuel to the air stream and break up the fuel into small droplets so it can be carried by that air stream. The resulting combustible mixture should be homogeneous and suitable for positive and economical Operation of the engine. As the air velocity past the jets increases, the fuel particles discharged from the jets become smaller and the proportion of large particles decreases, giving better mixture conditions at the carbu~ retor. The condition of the mixture at the cylinder depends upon the Shape of the manifold, the velocity of flow, and the heat energy imparted to the mixture. The intake manifold should distribute the carburetor mixture to all cylinders in equal proportions or maintain the same mass flow to all cylinders. There is not a manifold designed which can accomplish this throughout the range of Oper— ating conditions. The following are some of the conditions or restrictions placed on the design of an induction system. Relative engine load is one of the most important factors affecting the performance of the intake manifold. Dilution of the 3 fresh charge by residual exhaust gas varies inversely with engine load. For example, at constant Speed when Operating at high load, there is more fuel-air mixture inducted per cycle into each cylinder than at low load. The degree of dilution in each cylinder varies from cycle to cycle as the exhaust pressure in each cylinder causing the dilution varies from cycle to cycle. Therefore, it is almost impossible to induct equal quantities of fuel and air into each cylinder. The particular purpose for which the engine is designed has an effect. High-output engines require large flow areas and a fine diSpersion of the fuel in the mixture or the so-called dry-mix. Engines designed for good economy can use small flow areas with their resulting higher velocities which will support a wet-mix. The requirement of a wet mix is that the quantity of fuel admitted to the air stream must be such that at the end of compression, all fuel is vaporized. The firing order has a large effect on the process of induc- tion. Too many cylinders drawing mixture from the Same branch successively promote unequal distribution, particularly at part throttle and light load. Reversals of flow in branches cause drop- Out of fuel droplets from the mixture. I By trial and error, it has q l’l.l 5 been established that not more than three cylinders may fire in any one branch without materially affecting the distribution of charge. Manifolds are designed for a minimum of reversals. The clearance gases flash back into the intake manifold upon opening of the intake valve, and cause a reversal of flow. At the same time, if the exhaust manifold pressure is high, it will irrlpede the flow of mixture. Valve overlap has an important rela- tion in this respect. The higher the valve overlap, the more the re is a chance for blow-back into the manifold. This flash-back or b1~0W-back dilutes the incoming charge, or if it does not return to the same cylinder, it will, cause a rich mixture there, and a lean mixture elsewhere. Once the cylinder pressure is reduced below 1‘-hat of the manifold the flow of mixture will start. Available time fOr charging of any cylinder is lessened by any factor which Creates a resistance to the mixture flow. Therefore, less charge Can enter that cylinder. A circular sectional area will have the least flow resistance, but minimum surface area is not always a desirable feature. If the mixture is wet, heating of the mixture might be desirable so that the precipitated liquid fuel can be returned to the mix. En- trainment of the fuel is never complete, so some provision should be made to aid the return of the precipitated particles to the mix— ture _ The separation of the mixture in the manifold begins at the manifold tee. Because of the inertia of heavy particles, a puddle 0f fuel will form at the base of the tee, and this section should be level and perpendicular to the flow axis, so that the liquid fuel will not flow more into one branch than into the other, by gravity or inertia action. If this occurs, one branch will be rich and the other will be 1ean, indicating poor distribution. This perpendicular section is uSually heated by the exhaust gas so that a limitation is imposed upon the amount of fuel than can collect. This also aids in making the mixture more homogeneous. The section below the tee may be a depression in the manifold so that the forward motion of the Vehicle will not cause the liquid fuel to flow to the rear branch by inertia. This rear branch can be elevated slightly to limit the IiQuid flow. This applies to long straight branches of a manifold that are found in large in-line engines. The throttle valve prohibits a uniform mixture in the riser seCtion, eSpecially at part—throttle operation. The fuel particles Will be deflected from the throttle valve to one side of the riser and will tend to flow into the branch connected at that point. This enrichens the cylinders fed by this branch. When this occurs, the best condition of mixture distribution for economy is to have one rich cylinder and the rest receive a uniform mixture. The poorest condition for economy would be the inverse, where one cylinder is lean and the rest receive the same mixture. With one lean cylinder the mixture ratio must be enriched in order to fire the lean cylinder and the balance of the cylinders will receive excessive amounts of fuel. Maximum power mixture ratios with gasoline as fuel are in the vicinity of 12.5:1, where maximum economy mixture ratios Vary from 13:1 to 20:1. The combustible range in mixture ratio is about 7:1 to 20:1, dependent upon the particular engine character— istics. Some engines will run satisfactorily on a 16:1 or higher mixture ratio, while another of the same design will not fire the Same mixture continuously without missing. BRIEF STATEMENT OF THE PROBLEM The main object in this experiment is a study of the relation between intake-manifold branch pressure differential and indicated horsepower. Differential pressures are measured for each cylin- der between the base of the carburetor and the intake port. Indi- Cated horsepower is recorded for the same operating conditions. In testing, the throttle was fixed and increments in r.p.m. were Varied by a change in load. Four manifolds were tested so that the conclusions will not be judged by the Operation of any one intake manifold. All data recorded in the Appendix are not the aVerage data, but the values which were the most consistent. Performance characteristics of the engine for each intake manifold iIlstallation may help to distinguish the most effective induction System with particular emphasis on output. CONCLUSIONS There is not a specific variation or a definite relationship between pressure differential and developed engine horsepower. There are other variables present that affect the state of distribu- tion. Load, r.p.m. , and flow rate are the most important of these. The developed horsepower will be more uniform when the load and r.p.m. are high. At low r.p.m. and high load, the develOped horsepower will be higher per cylinder, if the pressure differential is high with the low-pressure end of the differential 1~0cated at the intake port. This does not imply that the develOped horsepower will only be a maximum under these conditions. The Same high horsepower can be obtained with a low-pressure differ- ential with the valve having the higher pressure. This is most likely due to the liquid end of the fuel that enters that cylinder at 1111at time. In other words, when ramming effect is present, the develOped horsepower will be high. When ramming effect is ab- Sent, the horsepower is usually lower. At high flow rates, the pressure differential has little effect On the develOped horsepower. Under conditions of high flow rates, the ramming effect is allnost always present. When it is not, 9 all. 10 the absolute pressure is high enough to obtain an adequate flow to the cylinde r. The load factor Shifts the point where ram initiates. As the load is increased, the point where ram starts is reduced. For an example, Operating at 1/4 throttle Opening and 2000 r.p.m. , the point where ram starts might be 2500 r.p.m. Upon Opening the throttle for higher load and holding the r.p.m. at 2000, the point Where ram starts will be lower than 2500 r.p.m. The liquid—fuel particles do not have a definite flow pattern. Under assumed low volumetric efficiency conditions, the minimum Velocity in the largest intake manifold branch is Slightly less than 1-0 ft. /sec. The minimum carrying velocity for a liquid fuel of the Specific gravity of gasoline is 2 ft. /sec. From this, drop-out 0f the liquid end will not occur except by an instantaneous zero Velocity from reversals of flow or intermittent flow. The control of liquid fuel is of major importance in obtaining high brake mean effective pressure and volumetric efficiency curves. Performance curves were drawn for all conditions of opera- tion. The only curves that can be compared are the full throttle Curves, as the load varied from one manifold to another at the Same part-throttle opening. These curves can be adjusted for 11 comparative performance, but the pressure differential cannot be adjusted. The data for pressure differential must be taken under Operating conditions. The full-throttle brake horsepower curves for all induction Systems fall within a 1 percent variation for engine Speeds below 3800 r.p.m. Further increases in r.p.m. exhibit a little more variation. The brake specific fuel conSMption varied between manifolds. The miles per gallon would not be a measuring device as each manifold has its own advantages, dependent upon the range of r.p.m. desired. In comparing the friction horsepower curves of all manifolds at full throttle, the dual-carburetor manifold was expected to be the least. This. is not true; the larger cross—sectional area manifold, the quad, was the lowest in frictional horsepower. One could as- sume from this that cross-sectional area and internal surface are more important to frictional resistance than the length of flow Path, at least in the testing range of this experiment. This might not be true at higher r.p.m. Performance is higher in the quad. manifold when operating at 3/4 throttle, than when operating at full throttle. Throttling of mixture must occur at the intake valve or the displacement is not high enough to cause higher flow rates. 12 The quad. carburetor adapted to the standard manifold had no apparent effect on the performance of the engine. The fuel consumption was a little higher, but the output compared to the rest of the manifolds. The frictional horsepower was the same as the standard manifold and carburetor. In brief summary, under an absolute pressure analysis, if the absolute pressure is high near the valve, the develOped horse— Power will be high. ii DISCUSSION OF PROBLEM The particular problem in this experiment is the relation- Ship between the differential pressure and the developed horse- power of each cylinder. The differential pressure is that change in pressure that occurs between the base of the carburetor and the intake port. This is completely within the limits of the induction System. This is not a discussion of the mixture ratio produced by the carburetor, except to assume the carburetor will meter the fuel uniformly, regardless of the actual mixture ratio. This is a discussion of the distribution of the mixture to each cylinder and the determination of the influence of pressure change on de— Ve10ped horsepower. The mixture ratio in each cylinder is not measured, but if the develOped horsepower of each cylinder is equal to that of the other cylinders, one might assume their mix— ture ratios are similar. The pressure differential could reflect this condition. The Optimum condition in the operation of an induction System will be when the maximum possible weight of mixture passes the intake valve. The mixture should be as close to 13 14 atmospheric pressure and temperature as possible. When any fluid flows through a closed channel, there is always some pressure loss caused by friction against the wall surface and a pressure loss by shear of adjacent layers of fluid which move at different veloc— ities. The Reynolds number is always sufficiently high so that Streamline flow is almost an impossible situation. Operation is almost entirely in the turbulent range of the Reynolds number, but at very low r.p.m. , the intermediate condition is sometimes obtained. When operation is within this region, the flow is pulsating so that all the variables present are not constant. Any Imbalance will tend toward the turbulent condition which is the usual state. Therefore, a pressure loss from turbulence within the fluid caused by irregular, nonstreamline flow is a factor which will limit the weight of charge that can enter any cylinder. These pressure losses appear as heat in the mixture. The magnitude of the energy losses are dependent upon the type of fluid, the relative roughness of the flow path, the configuration and length of the flow path, the cross-sectional area of the path, and the rate of flow. There are too many variables present to be able to design a manifold for certain Operational conditions. The l5 effect of wall roughness and configuration cannot be predetermined; 131111 for minimum pressure losses, keep all bends as streamlined as Possible and the channel walls smooth. Pressure losses due to rate of flow, length of flow path, and the cross—sectional area can be calculated as the pressure losses increase with the square 0f the rate of flow. Small cross-sectional area manifolds will have high velocity of flow 1ill—roughout the range of r.p.m. Drop—out of entrained fuel will be less when pressures change by throttle control. Engine output at high r.p.m. is limited by the throttling effect of the Small cross-sectional manifold. In order for some of the past variables to enter into this problem, a dual carburetor manifold was. chosen because of its different flow conditions, such as length of flow path and the rate of flow in that path. The cross-sectional areas are the same as in the standard production manifold, but the pressure losses should be one-fourth of those in the standard manifold. This is approxi- rnettely correct because the rate of flow is reduced by the addition of another carburetor. If one can assume the rates of flow through each carburetor are equal, the pressure losses will be one—fourth 0f their original value with one carburetor. This in turn reduces 16 the mixture velocity and drop-out of liquid fuel will be higher at corresponding speeds. At very low speed, an over—carbureted con- dition may develOp upon sudden opening Of the throttle valves due t0 a 1a rge drOp-Out of liquid fuel by the sudden change in pressure in the intake manifold. At high r.p.m., there should be no loss of powar by the throttling effect of the small manifold as the flow rates are about one-half of that in the standard manifold. A manifold larger in cross—sectional area and a carburetor ’ from a large displacement engine called a quad. carburetor and manifold was tested. This carburetor is designed so that one throt‘l‘ole, designated as the primary throttle, controls the rate Of flow through the manifold until a particular opening of that throttle is I‘eached. At this time, another throttle called the secondary throttle starts to Open into the same branches fed by the primary. Both throttles reach the full Open position simultaneously. This I‘nahifold being larger in cross-sectional area should exhibit prOp- erties between the dual-carburetor manifold and the standard Irlanifold. It should be superior to the dual-carburetor manifold at low Speed due to the higher velocity Of flow. It should com- pal1‘t-':.'with the dual~carburetor manifold at high Speeds if throttling in the manifold does not occur. 17 In order to determine the relative position or the more critical component in the induction system, an adapter was made t0 fit me large quad. carburetor to a standard manifold Of smaller cross—sectional area than the carburetor. Even though the primary Pm'POSe of this investigation is to determine the effect of pressure differential on develOped engine horsepower, the performance of the above manifold and carburetor should give valuable information, eSPeCialJ.y if the maximum horsepower compares with. the rest of the inta ke manifolds. CHOICE OF OPERATING CONDITIONS The effect of differential pressures on the output of each cylinder is the prime Objective. Therefore, the Operating conditions should be such as to indicate a wide range in pressure differential. The brake mean effective pressure and the volumetric efficiency are aLit'fected at low Speeds by low flow rates, and these can be aggraVated by more blow-back in the intake manifold. A longer infflke-walve Opening duration, where the intake valves close later and Open sooner, Should exhibit more blow—back. At the same time, a 1C>I'iger overlap between intake and exhaust valves should give- poor low—Speed performance. These conditions Should increase the limits of pressure differential. The compression ratio was increased to insure complete VaPorization of the mixture upon compression by the correSponding inCrease in compression temperature and pressure. Variable Speed testing appeared to be more suitable with this tYpe of installation than constant Speed. The throttle Opening was held to a predetermined value and the r.p.m. was controlled by a change in load . 18 19 An eddy—current dynamometer was used for absorbtion of torque and the indicated horsepower was determined by the Short circuiting Of one cylinder and then summed for the total. The indicated horsepower of one cylinder was determined by the differ- ence in brake horsepower at the same Speed when all cylinders were fi‘ring and when one cylinder was shorted. The result was the inClicated horsepower Of the shorted cylinder. The friction hors’epower, determined by the difference between indicated and brake horsepowers, will be close to the actual value as engine operation more nearly approaches the firing conditions of all CylhIders. Four manifolds were tested with the following throttle Opening: Standard Manifold: 1/8, 1/4, 1/2, 3/4, and full throttle; Dual-Carburetor Manifold: 1/16, 1/3, 1/4, 1/2, and full 11hrOttle; Quad. Manifold with the secondary throttle Opening at 1/2 13l‘imary Throttle: 1/8, 1/4, 1/2, 3/4, and full throttle; Quad. Carburetor adapted to a standard manifold with the sEtcondary throttle Opening at 1/4 Primary Throttle: 1/8, 1/4, 1/ 2, and full throttle. For additional information on the quad. manifold and car- buretor, the secondary throttle was adjusted to Open at 1/4 primary 20 and data taken for the following throttle Openings: 1/8, 1/4, 1/2, and full throttle. The throttles of the dual carburetors were synchronized so both would open simultaneously. The throttle positions started at 1/16 throttle Opening in the hOpe of coming close to the same load as that obtained when the throttles Of the other carburetors were Open at 1/ 8 throttle. Another test with the dual carburetors would be to operate the throttle of the first carburetor while the second is in idle poshLion and then let the throttle of the second carburetor open to the same position as the first when Speed Of maximum torque is reached. This procedure should reduce the overcarbureted con- dition that occurs when both carburetor throttles are synchronized. When the large quad. carburetor was installed On a small displacement engine, the velocity Of flow through the carburetor Would be lower and a lean mixture could result, eSpecially at low 1“ P.m. Due to larger metering jets, the mixture may have been ml the rich side. Perhaps some of the above conditions or hypotheses could be proven from the pressure analysis and the performance char- aecteristics. PROCEDURE In Order to limit the number of variables that might enter into this project from the operation of the engine itself, the engine was completely rebuilt and the following components replaced or reconditioned. Rabore to 0. 080 inch oversize. New pistons and chrome rings (4 ring piston). New crankshaft and bearing inserts. New wrist pins and pin bushings. New timing gears and reground camshaft. New valve Springs and valve guidesg. Valves reconditioned. Valve seats reconditioned. Connecting rods checked and aligned. The clearance volume of each cylinder was checked so that the C1eveloped horsepower of each cylinder would be compared. By caIch'lation, if 0. 003 ”inch was removed from one cylinder head, the. CleaZl‘ance volumes would be very close. For an increase in com- DreSeion ratio, 0.045 inch was machined from one head and 0.042 inch from the other. The resulting clearance volumes were the 21 “*1! ‘. fl ‘0! ‘x‘r' :- 22 same for six cylinders and varied one ml. above and two ml. below for the other two cylinders. This little variation should not alter the compression ratio in each cylinder appreciably. The error would be less than 0.5 percent if the measure— ments were close. Before assembly of the engine, the intake ports were polished, then bored and tapped for tubing connections. These are the static pressure taps and were located as near to the intake valve as possible, but in such a position on the outside of a long radius bend that they were not near a stagnation point or an eddy-current center. These pressure taps were drilled for an 0. OBI-inch orifice to limit the flow and fluctuation of the liquid in the U tube manometers. The engine was connected to the dynamometer and the dynamometer was corrected and adjusted for balance. The scale reading was twice the actual applied force. The standard intake manifold was tapped for static pressure measurements in each branch of the riser section and as close to the base of the carburetor as possible. The change in the direc- tion of the airstream by the throttle valve could cause a velocity pressure at the orifice of the [pressure tap. Thus, the static pressure taps were located at the side of the throttle valve to Z3 eliminate direct impingement of the airstream on the pressure tap. These pressure taps also were drilled for a 0.051-inch orifice, but later changed to a larger size to reduce the loss of measuring liquid in the manometers by sudden changes in static pressure from acceleration of the engine. All manifolds were equipped in the same manner, although orifice size varied in some. In order to obtain data quickly and efficiently, the perform- ance tests were run first where atmospheric conditions have a great influence on final results. All correction factors were according to S.A.E. Specifications. The following data were recorded for the performance tests: Throttle Opening r.p.m. Beam Load-«lbs. Pressure Differential--"HZO (C and V refer to the point of lower pressure.) Cylinder Numbe r - mWWNNO‘rP-I-i Oil Pressure ——#/" 24 Jacket Temp. --°F Intake Vac. --"Hg S. P. Adv. -—°BTDC Exhaust Press. , L. B. --"Hg R. B. --"Hg Weight of Fuel--oz. Revolutions-- Time of Rev.--Min. Fuel Temp. --°F Dry Bulb Temp. --°F Wet Bulb Temp. --°F Bar. Press. -—-"Hg Some of the data were not used in this experiment, but might prove useful for someone who might have another approach to this problem or a similar project. In this problem, atmOSpheric conditions do not have an effect on the determination of indicated horsepower per cylinder as a comparison or trend between the pressure differential and indicated horsepower per cylinder as a comparison or trend between the pressure differential and indicated horsepower is the object. It is not necessary to use a correction factor on the indicated horse- power per cylinder. 25 Further-data were taken for the evaluation of friction horse- power and indicated horsepower per cylinder, and the classification follows: Throttle Opening--same as performance tests. r.p.m. --same as performance tests. Beam Load-~all cylinders firing---# Beam Load--#l cylinder grounded---# Beam Load--#2 cylinder grounded---# Beam Load-«#3 cylinder grounded---# Beam Load--#4 cylinder grounded---# Beam Load--#5 cylinder grounded-—-# Beam Load--#6 cylinder grounded---# Beam Load--#7 cylinder grounded---# Beam Load—-—#8 cylinder grounded---# Beam Load—«all cylinders firing---—#. DESCRIPTION OF APPARATUS The absorption unit. was a Midwest Dynamatic Dynamometer of the eddy-current type. It had a water-cooled field with a heat exchanger and circulating pump. The horsepower rating was 175, and the Operating constants were as follows: iBeann ggadjfrm. m.) 8000 HP = Torque = (0.6565)(Beam Load) in ft. # The torque arm is 15.756 inches. The electrical input to the dynamometer is supplied from either a D. C. line or the control panel where close regulation is possible. The. r.p.m. indicator, the revolution counter, and the clock are a synchronized assembly, Operating from the dynamometer and an electrical source. The jacket temperature of the engine was controlled by a heat exchanger where the flow of cooling water is manually Operated. An automobile radiator was submerged in a tank and the cooling Water flowed around the core of the radiator. The exhaust system had an external water jacket, but this Was only a safety feature to guard against burning to anyone who might touch it. 26 DESCRIPTION OF TEST RESULTS Discussion of Figures 1-22 This section is not intended to credit or discredit the per- formance of any particular induction system, but only to describe and perhaps clarify any inconsistencies that might appear in each system. Due to the radical change in valve timing, some of the results good or bad might be assessed to the induction system, when they are actually an internal effect. Each curve will not be dis- cussed. Only those curves that have peculiar shapes will be dis— cussed with a possible explanation to each. Some of the curves do not apply directly to this investigation, but are presented so that anyone who is interested in another phase of investigation may have the data to work from. A complete listing of the curves ap— pears in the List of Figures. These figures are arranged in numerical order and appear at the end of this section. The range of r.p.m. was not sufficiently high to cause an intersection of some curves or a definite drop in performance. In Figure 1, the no-load Speed is approximately 1850 r.p.m. , where an intersection of the friction horsepower and indicated 27 28 horsepower curves will occur and the torque and brake horsepower curves simultaneously go to zero. This is not shown in Figure 1, as the maximum r.p.m. reading is 1600. In Figure 2, the indicated horsepower curve has a definite dip at 2000 r.p.m. This is due to a rotational inertia unbalance. At the time of testing, a pronounced roughness at 1050 r.p.m. was exhibited and the same condition was noticeable at 2100 r.p.m., but the amplitude of vibration was not as high. The indicated horsepower and torque curves have a dip in Figure 3, and this is reflected on the brake—horsepower curve as a flat Spot, all in the range of 3000 to 3500 r.p.m. A roughness of Operation was not apparent, but might have been dampened due to the large flywheel effect caused by the heavy stator in the dynamometer. The torque curve of Figure 5 has. a very definite dip at 1000 r.p.m. In Figures 3 and 4, this is not apparent as the initial reading was taken at 1000 r.p.m. where the dip should occur. If the initial reading was 500 r.p.m. as in Figure 5, more definite conclusions could be drawn at the 1000 r.p.m. reading. Inertia unbalance is the probable cause for the loss of horsepower and torque . . .11 1|ll Ill. ‘ l l 29 The torque and indicated horsepower curves exhibit a slight loss of performance in the range of r.p.m. from 1500 to 2000, Figure 6. This loss is reflected in Figure 7, but the r.p.m. range is a little higher. All of the curves in Figures 8, 9, and 10 are smooth except the indicated horsepower curves in Figures 9 and 10 around 3600 r.p. m - This was probably due to the high frictional horsepower at this speed. A small percentage of error in the calculations in- volving the indicated horsepower for each cylinder will account for a larger percentage of error in the corrected indicated horsepower tot-211$. Larger values are obtained at higher r.p.m. , as the indi- cated horsepower varies directly with the r.p.m. , at least within 10 pe rcent of maximum r.p.m. If these indicated horsepower dips were caused by harnnonic vibrations, these vibrations were small in rrla~gnitude and the lower harmonics do not appear in the 185111155 Where the amplitudes were necessarily much higher. The exhaust back pressure suddenly increased at this point (no longer a linear variation with r.p.m.), but for that increase the output was not reduced to any great extent. This particular condition must have bean a combination of factors, at least more evidence is necessary {01' a more definite explanation. 30 The performance curves for the quad. manifold covering Figures 11 to 15 have few ranges that are not smooth. In Figure 12, the torque and indicated horsepower curves have slght dips at 2000 r-p.m. Figures 14 and 15, the dips occur on torque, indi- cated horsepower and brake horsepower curves at 3000 r.p.m. The reason again was inertia unbalance where the higher harmonics gave rise to the loss of horsepower and torque. Figures 16 to 19 refer to the special manifold where the quad. carburetor was adapted to a standard manifold. In Figure 17 at 2-000 r.p.m. , the indicated horsepower and torque curves show a decrease in performance and the same is true in Figure 18, but at 3100 r.p.m. with the addition of the brake horsepower curve and its flat section. This was due to inertia unbalance. Figures 20, 21, and 22 indicate the same conditions of opera- tion as the previous figures except for the brake horsepower curve 1 n Figure 22, where a flat spot covers a wider range in r.p.m. , f “In 2300- to 4200. At the time of testing there were several possible explana— t‘ . tons for the configurations of the curves just discussed. Some merited further investigation, but most had some disadvantages. 301116: of these theories follow. 31 Torsional vibration of the crankshaft could be the cause where the second harmonic is approximately 900 r.p.m. , the fourth harmonic at 1800 r.p.m., and the sixth harmonic at 2700 r.p.m. If this were true, the vibration would occur at these r.p.m. read- ings regardless of throttle position and load. The resonant condi- tion seemed to vary with load. When the load and r.p.m. readings were high with wider throttle Openings, the lower r.p.m. ranges with the high load did not exhibit the same magnitude of vibration as the light load condition. If the vibration was present there must have been an unknown damping factor which limited the magnitude or Vibration. At very light loads the point of resonance was at 10‘” r, p.m. and was not present at high r.p.m. The intermediate condition of load and throttle position may exhibit resonance at mid r.p.m. range and again in the high r.p.m. range. The low r.p.m. range seldom showed a resonant condition. Another possible solution was column vibrations from pGui-Odie rebounds in the long exhaust system. This should give a sudden increase in exhaust back pressure by reinforcement of the Sound waves. The data on exhaust pressure does not support this theory at all resonant conditions. The intake manifold might have had column vibrations but the“! Would have been very short in wave length. The time element 32 for the rebounds of the sound waves would be very short to travel from the intake valve to the carburetor and back again when the intake valve is in the exact Opposite position in. the succeeding cycle. AtmOSpheric conditions and variable charging pressures should vary the results from day to day. This is not apparent from the data. For the above condition any change of any variable would upset the timing Of the sequence and the column vibrations would not annihilate each other. A reinforcement Of these waves would cause a gain in horsepower from a psdueo-supercharger. Due to blow-back in the manifold at low r.p.m. , the mixture ratio might have been the variable giving low performance from improper mixture. The higher r.p.m. ranges might have been a transient condition in mixture ratio or poor carburetor char- aeteristics. This cannot be true, as the various manifolds had flat Spots or dips at the same r.p.m. changes. The products 0f Ccmnblustion were analyzed by an engine analyzer and the mixture ratio Was always within the combustible range. After testing was completed, all rotating masses were balanced and this mmoved the roughness at all engine speeds except the 1000 r.p.m. condition. At this point, the magnitude very small in comparison tO the roughness at the time Of 33 testing - Number seven piston was heavy. Holes were drilled in the skirt so that all pistons were of the same weight. Discussiongof Fijrures 23—56 A description or discussion Of the curves in this section is itnpra Ctical as one curve in itself will not support any conclusions that might be drawn from the trend Of pressure differential. In- Stead: these curves have value as reflections Of other curves that are not drawn. The brake mean effective pressure curve indicates the gen- eral Shape Of the volumetric efficiency curve. These curves are “rho-fit an exact duplication Of one another at low and medium Speed. The breathing efficiency Of an engine is a direct measure Of the brake mean effective pressure if friction is not considered. The trend of mixture ratio is indicated by the brake mean ef- fectin pressure. If the mixture ratio is toward the lean, low °utput results or low brake mean effective pressure and high brake Specific fuel consumption is Obtained. The weight of the charge is reduced as the engine Speed incl‘eases because higher velocities are necessary to move more Charge to the cylinder. The higher velocities are caused by a 39 greater pressure drop in the induction system, resulting in lower cylinder pressures and therefore less weight Of charge. The brake mean effective pressure curves should relate these conditions. The pumping losses are part Of the total friction horsepower and tlle trend of the mechanical efficiency curves should indicate thB Comparisons in these losses. Data for demand horsepower curves were not recorded but miles per gallon curves were drawn for constant throttle Openings. vehicles operate at variable throttle, so one must not draw con— dusions concerning vehicle performance directly from the miles per gallon curves. Comparisons between manifolds at the same throttle opening must be tempered by the load factor, as the load varied between manifolds at the same throttle Opening and r.p.m. The curves in Figures 42 to 56 are intended for general in'fol‘l'nation. The comparison Of these curves shows the trends Of eaCh characteristic upon Opening and closing of the throttle. Dissatisfied of {Leases 51.200 Figures. 57 to 100 are arranged face to face with all Operating conditions the same so that comparisons or trends can be drawn. Each line graph is drawn according to individual 4O induction systems, and only refer to those cylinders fed by one riser. All intake manifolds have at least two separate branch Systems. The dual-carburetor manifold has four risers, but two small connecting passages join the risers in pairs. These graphs have reference to the change in pressure within the extremes of the branch systems. With higher velocities 0f flow, a. greater differential in pressure must exist between the atmosphere and the cylinder to accomplish adequate mixture flow in the short time available for charging. This is not measured by the internal pressure differential in which this problem is con- cerned . Standard Manifold Figures 5] and 584 11s throttle Opining. Cylinder number six has a higher than average indicated horsepower value and the pressure differential is less than average. Cylinder number five has a low develOped horsepower value and an average value Of pres- Sure differential. The points Of higher pressure occur at the valve. Fi es 5 a d 60 1 4 throttle 0 e . Cylinder number Six has a high indicated horsepower and a low pressure differential. CYiinder number four has a high indicated horsepower and a high 41 pressure differential. Higher pressures are located at the valve but are lower than previous figures. Figures 61 and 6;; 1/2 throttle Opening. There is not much fluctuation in the develOped horsepower except cylinder numbers six and seven are slightly higher than the rest. The pressure differential does not exhibit a very wide range, but cylinders six and Seven are a little lower than average- The Point Of higher pregame fluctuates from one end of the induction system to the Other. The ramming effect is just starting. Figures 63 and 644 3/4Vthroftvtleggggenipg. The graphs of developed horsepower exhibit little variance. Cylinder number seven is high at 4000 r.p.m. The pressure differential is high for cylinder seven with the valve pressure the lower. Cylinder four has an almost constant pressure, but is higher than average throughout the range of r.p.m. Almost all pressure measurements Show ram in good progress. Figures .65 and, 66, full Mottlgggenmg. The graphs are Corrlparable to Figures 63 and 64, except the change in pressure is more gradual between r.p.m. increments. The develOped 42 horsepower is higher and the pressure differential is higher. Ram is pre sent at all conditions of Operations. Dual—Carburetor Manifold Cylinder numbers 1, 6, Z, and 5 are fed by one carburetor, and 3, 8, 4, and 7 by the other carburetor. There are four risers and each feed two cylinders as follows: 1 and 6, 2 and 5, 3 and 8, 4: and 7, A small passage between risers connects the following CYIinde IS in the branch systems: One system is Z and 5, 3 and 8; the c“filler system is 1 and 6, 4 and 7. Figures 6]; and 68Lg1416 throttfilgtgpetning. A trend is not appal‘ent in these figures. There is too much variation in pres— Sure and develOped horsepower. Figures 69 and 794,118 throttlegpening. Cylinder number 51" has a comparatively low pressure differential and a low de— Velo'Ped horsepower. Cylinder number seven has a low horsepower and the lowest pressure differential and shows ram at low r.p.m. T1131 point of higher pressure is the valve end of the induction system. 43 Figures 7Land 7;,‘i[4 throttle ogggng. There is little variation between these figures and the previous figures. The develOped horsepower is higher and the valve pressure a little lower. Fjgfiufires 73 and 74, i442 throtflggpening. The developed horsepower is practically a constant for all cylinders. One riser section of one carburetor shows high valve pressures and the» other riser, low valve pressures. The other carburetor has the Same conditions. This must be due to flow resistance. Figures: 7454 andfl76, full throttle om, The operating conditions are the same as the 1/2 throttle position. Developed hora"‘kpower is slightly higher, but the variations are also higher. The increased flow rate shows greater losses in cylinders 1 and 6’ 3 and 8, but the output of all cylinders is not different. Quad. Manifold In this manifold the secondary throttles feed the same cYunders as the primary throttles. Figure} frigid 78,4113 throttle opening. Cylinder seven has a lower than average develOped horsepower and the pressure 44 differential is low. Cylinder number eight has high output and the pressure differential gradually decreases with increase in r. p. m. This cylinder has high differential at low r.p.m. with the valve end of the induction system being the‘greatest. The increase in r. p.m. is accompanied with a lower pressure differential until the pressure is the same at both ends of the induction system. There is some fluctuation in the point of maximum pressure, but generally, the valve end is the higher. Figures 79 and 89,:inthrofile opening. Cylinder number one has a low pressure differential and an average indicated horse— PoWer, Cylinder number four has an average output and a high Pressure differential. Cylinder number five has a low pressure diffe’I‘Vential and a high output. There is not definite point of maxitnum pressure, but varies between a range of 2" H20 dif- ferential around the 0" H20 differential point. Figures, £1 anL82,,1[g throttle opening. (This is the Position where the secondary throttles start to open.) One induc- tion system has a high pressure differential with the valve end the higher for all values and the other system a low pressure differ- ential with all values the lower at the valve end of the induction 45 system. The later condition exhibits a more uniform distribution of develOped horsepower and the pressure variation is very small around the 0" H20 value of pressure differential. Figures 83 and 843 3/4 thfirottle opening. Most of the values of pressure differential are fairly uniform for all cylinders and the valve end of the induction system is the lower. The develOped horsepower is almost constant for all cylinders. Figuggg g8,5 andv 86, full throttle Opening. The pressure differentiai changed slightly from the 3/4 throttle opening. The developed horsepowers are lower at low r.p.m. and have identical valves at higher r.p.m. There is evidence of throttling in the 1“duction sy stem . Quad. Carburetor Adapted to a Standard Manifold (the secondary throttle open at 1/4 primary throttle) Figures 37 and, 88,}[8 throttle Opening. Distribution of PNSSure differential is uniform and limited over a small range. In all cases the valve end of the induction system is the higher Pressure. Cylinders one arxi four have low values of pressure differential and the develOped horsepower is higher than average. 46 Figures 89 andViO, 1/4 thrfiottle Opening. There is a distinct variation in developed horsepower but there is no apparent direction or trend in the pressure differential as there is extreme fluctuation. The average pressure differential is high with the valve end of the induction system the higher. Figures 91 and ‘32,le throttle opening. There is a uniform pressure variation with all values of pressure differential lower at the valves. The developed horsepower is regular with very good distribution, although cylinders seven and two have the lowest pres- sures. Ramming effect is present. Figures 913 and 94gfulLthQ§le Opening. The only differ- ences between these figures and the past two figures are the ex— tremes of variation. The Output is higher and the range Of pressure diff-erential variation is greater, although the average numerical value 15 lower. Ramming effect is present. Quad. Manifold (secondary throttles Open at 1/4 primary throttles) Figures 25 and 26, 1(4 throttle Opening. The pressure differential is evenly distributed for all cylinders. The developed horsepower is higher than average, but cylinder number seven is 47 low with no pressure differential. Cylinder number four has a higher differential and an average output. Ram is present in some cylinde rs. Eigures 97 and 9,8: IZLthrgttle opening. One induction SYStern feeding cylinders 2, 3, 5, and 8 has lower valves of pres- sure differential than the other induction system that feeds cyl- inde rs 1, 4, 6, and 7, The output of the former is higher than the Output of the latter. Figure}; 99 and 100$qu throttle opening. The pressure diffe rential is similar to that of the previous figures. The indi» Gated horsepower is higher but regular from cylinder to cylinder. PERFORMANCE CURVES Figure 5 1-22 Indicated Horsepower Brake Horsepower Friction Horsepower Torque 48 49 \ OE {mm once .0: no: ogu no? IT - u d [A O m u. l: dzH can. on 3 3 936er m: 2.3V 81 ur {revenue QQO\~\<d(t.um «. +( «219 xtod4+ xu<~so¥w 02.321 AM. 04454: .330 n usuaU Uuidtu 0¥N um coax cahx .Z .0 M 002 DU? .3 £\1 o~ 2. 83 KI‘I.QQ pm *1 4.23.0 u‘tbu.:\ Xatrnouow O V it. «4+ ..I O 4 g». .IC—L .flCJG mgusU 0921... OOWUVO— .: 3: —\ fin .@ «It 0.0M .00“ 08_ .z a .m m 0% Ar ¢ m * / .Imv .I o: a 90‘ Hanna _ .1 _ u .t .ui «.85 - .. fl 34 .2 .U.‘ .m... . 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My, 'Ie. ; _ ,_ i , ,,,,,_,_____ I000 37.3'3-1313; 46);.JFQ-fwwfi-f ' V-J' 431: $37! 43" 73‘. 2-3' Ira. «4.» 4w 7.» 1..» 7.31. L“! 6.74 4'71 amt/75"»!- M86 20::- bm’ 1.34: (3.»: 1.»: 4.»: 34.»: 5.»! 9.0 (a! 73.1: m: use 76, Mr Mir1 Mr (“7' 11"“ ”4'! ,, 9""? 3":‘{-+13;33.- 13:3 3.0;: 84:. D3.“ liffLBSS '3 ‘5’}? 5,5 7359'.,L,‘.33,§§'_,,_‘3'7.57”0'3 ”Ll. 350:: j}. T1&7; 57:]er _LSJIJIL‘IJ' NW at)! (LTIJ’Jf‘LH3F‘g _3 4%» __W:_,,,_z_7._, 1’7. ,L/Jo 7’77””. ,/7.,,,n. .136. 42- L , _,,.______ J . _,_ - _L _-_, _ 1,...L-m,,.i - L APPENDIX C Engine Dimensions and Tolerance Assumed Constants 189 Engine APPENDIX C Dimensions and Tolerange Assumgd Constants C. R. 7. 15 - l Bore 3.2675 in. :1: .0004 in. Stroke 3. 75 in. Displacement 255 cu. in. cl. vol. 41.5 cu. in. Crankshaft Mains 2.499 in. (. 001 in. -.002 in.) Rods 2. 139 in. (.0005 in. -.0015 in.) Valve Clea rance s Intake . 009 in. Exhaust .011 in. Valve Spring Pressure 5 Valve Seat Angles Piston Skirt Clearances Area of Head Gasket Thickness Vol. . 945 42-45# at 1. 890 in. 52-55# at 1. 790 in. 45° .001 in. :1: .0005 in. 15. 1 sq. in. .0625 in. when compressed cu. in. 190 . 'i-l ‘f’ 1-"". 191 Clearances in Cylinder Head Cylinde r Numbe r 1 2 3 4 5 6 7 8 A bove .125.. .125.. .125" .125" .125" .125" .125" .125” Piston “me .156" .156" .156" .171” .1094" .1875" .1875" .141" Ex. Valve A b°Ve .156" .156" .156" .156" .141" .171" .141" .1875" " In. Valve . 1 _ 174:1 n 35 85 85 85 85 83 86 85 Minimum clearance - Ex. Valve of cylinder 5 - .1094". Standard Camshaft Ijgw Camshaft 1.0. 0° T.D.C. I.O. 20° B T D.C. Ex. C. 6° A.T.D. C. Ex. C. 18° A.T D.C. I. C. 44° A.B.D.C. I. C. 64° A.B.D.C. Ex. 0. 48° B.B.D.C. Ex. 0. 66° B.B D.C. In. Dur . 224 ° In. Dur. 264 ° Ex. Dur. 234° Ex. Dur. 264° Overlap 6 ° Overlap 38° Constants for Calculations: Mean tire rolling radius 14. 5" Rear axle ratio 3. 54:1 APPENDIX D Selected Refe rences 192 SELECTED REFERENCES MulticylinderiEngintefi‘Dgtonation Land Mixture LDistributfiwonJ Blackwood, Kass and Lewis. S..A.VEL gournaL March, 1939. Induction~ Maniolding, Louis Mantell, Automobile Engineer, July, 1940. Internal Corribgtign§g&Me§, Lichty, 6th Edition. Internal ombustion E ines, Polson, 2nd Edition. flh Speed @bpstion TEngines, Heldt, 14th Edition. Heat Transferavndv Eluvid'Flglv, Brown and Mares. Elementail LMve cpavnicsfi of Ft‘hiid sJ Rouse , 193 APPENDIX E Picture 5 194 Standard Intake Manifold “--.—r 195 Dual Carburetor Intake Manifold 196 Quad. Carburetor Intake Manifold 197 Quad. Carburetor Adapted to a Standard Intake Manifold 198 (left side) Engine Installation and Testing Equipment Engine Installation and Testing Equipment 200 (right side) 201 Exit of Pressure Tubes from Intake Ports 202 Engine and Accessories 1' ... '. 204 R. P.M. Indicator and Synchronous Clock with Revolution Counter I I: S‘Cl _ -__ ' L Engine Installation and Testing Equipment (left side) 205 ,1; 1 . 1 ‘ ‘ -' 3": K! in," ’t' 'f\-"_. n" at I: USE GMY