THE UNIVERSITY OF MICHIGAN COLLEGE OF ENGINEERING Department of Mechanical Engineering Technical Report ENGINE PERFORMANCE WITH VARIOUS COMBINATIONS OF DIRECT DRIVE SUPERCHARGERS AND TURBOCHARGERS E. T. Vincent N. Tokuda ORA'Project 058-47 under contract with: U. S. ARMY DETROIT PROCUREMENT DISTRICT CONTRACT NO. DA-20-018-AMC-0729-T DETROIT, MICHIGAN administered through: OFFICE OF RESEARCH ADMINISTRATION ANN ARBOR December 1964

TABLE OF CONTENTS Page LIST OF TABLES v LIST OF FIGURES vii OBJECT ix ABSTRACT xi INTRODUCTION 1 I. THE MECHANICALLY DRIVEN SUPERCHARGER ARRANGED FOR RESPONSIVENESS 3 A. Method of Calculation 4 B. Calculated Results 9 C. Responsiveness 14 D. Conclusions 23 II. COMBINATIONS OF THE MECHANICALLY DRIVEN SUPERCHARGER WITH TURBOCHARGERS 25 A. Displacement and TurbOchargers. Combined:: 25 B. Conclusions 25 C. Observations 28 REFERENCES 29 iii

LIST OF TABLES Table Page Io Full-Throttle Performance with Slip, Manifold Temperature = 200~F (F/A = 0.047) 10 IIo Part-Throttle Performance with Slip 12 III. Part-Throttle Performance, No Slip 13 IVo Responsive Performance with Two-Speed Fluid Drive Displacement Compressor 19 V. Performance at 3000 RPM 26 VI. Turbocharged Engine 26 v

LIST OF FIGURES Figure Page lo Diagrammatic arrangement of variable-ratio supercharger drive. 3 2o Typical performance map of a displacement supercharger with compression. 6 35 Supercharger characteristics plotted on a percentage basis. 8 4o Performance with slip at full- throttle. 11 5. Performance with fixed drive to supercharger. 15 6. Comparison of fixed and slipping drive performance. 16 7. Supercharger drive for responsiveness. 17 8. Response operation at full throttle. 21 9. Part-load responsive performance. 22 10. Full-throttle performance at 3000 rpm of combined chargers relative to a turbocharger alone. 27 vii

OBJECT The object of this analysis is to establish the expected engine performance when both engine-driven superchargers and turbochargers are combined in series, and to extend the performance data of Ref. 1 when the supercharger is driven in ways which permit some responsiveness to be achieved. ix

ABSTRACT The purpose of this report was to examine the ways and means by which the disadvantages of the turbocharged engine (low maximum power output at low speeds plus smoke during acceleration) might be overcome by combining it with the displacement compressor, and to examine the means by which the partload economy of a direct-drive, positive-displacement machine could be improved without sacrificing its high torque at low speed. Means were also examined by which some additional degree of responsiveness could be obtained with the displacement machine without sacrifice at all other load conditions. The results obtained show that the above objectives can be achieved, but that the turbocharged engine still gives the best overall fuel economy. Secondly, responsiveness plus increased economy at part-load can be achieved when using the positive-displacement supercharger, provided means are incorporated for variable-speed operation relative to the engine of the supercharger. xi

INTRODUCTION The methods employed to obtain the data for this report are those given in Refs. 1 and 2. Examples of the application of these methods are given in those references and the details of the calculations will not be repeated here, only the final results will be recorded. The report will be divided into two sections: Section I. The mechanically driven supercharger arranged for responsiveness; Section II. Combinations of the mechanically driven supercharger with turbochargers. 1

I. THE MECHANICALLY DRIVEN SUPERCHARGER ARRANGED FOR RESPONSIVENESS In this section a comparison will be made between engines fitted with a compression type displacement machine, such as the Bicera and Lysholm, driven in the following manner: Scheme (a) Direct constant-ratio drive Scheme (b) Direct variable-ratio drive (by fluid coupling) Scheme (c) Direct responsive-ratio drive (by fluid coupling). Fluid CouplingSupercharger7 Engine Speed Increaser Fig. 1. Diagrammatic arrangement of variable-ratio supercharger drive. Figure 1 is a diagram of the scheme. It consists of an engine driving the supercharger through any desired step-up ratio, which in turn drives one element of a fluid coupling, the second element of which drives the compression displacement supercharger. With this arrangement, Scheme (a) can be obtained by locking the fluid coupling, while (b) can be achieved by varying the amount of oil in the coupling; Scheme (c) also can be obtained by a variable amount of oil, but in a different manner from that of (b), as will be seen later. This type of'drive has been selected because it is simple, small, and light weight (when coupled to the high-speed shaft); it is well known 5

because of the years of experience with it in similar engine drives and its maintenance requirements are also well known. As a result of these factors, little development work will be requiredo The arrangement is, of course, still subject to the limitations given in Ref. 1 for this type of supercharger, viZo, 1o Relatively low isentropic efficiency 2o Relatively low maximum speed of rotation 35 Pressure ratios not exceeding 2.0 to 3.0:1 approximately, A.o METHOD OF CALCULATION In Scheme (a) the gear box ratio is selected to give the required speed of rotation of the supercharger when the fluid coupling is locked. For an engine speed of say 3000 rpm, a 3 to 4:1 ratio would be about the highest employed if the best of the supercharger characteristics are to be achieved. For the same reason, when employing Scheme (b) or (c) the same ratio of 3 to 4:1 would be necessary for the peak performance at maximum rpm of the engine. At part load, however, in Scheme (b) the fluid coupling would be allowed to slip in order to reduce the engine load at any given engine speed, i.e., road speed of the vehicle in any given gear. This means a reduction of the power required to drive the compressor, but also some loss due to slippage of the coupling; however the total power delivered by the engine to the supercharger will be reduced below that of Scheme (a) as the engine load is reduced for any given engine speedo The coupling's slippage, which increases as load is uced, means a reducing manifold pressure in place of a substantially constant one, as occurs in the case of direct driveo The result is not only reduced work of compression but some lowering of frictional losses also. In the operation according to Scheme (c) the reverse would be the case, slippage will exist at full speed and load so that the additional losses due to slip are encountered for short periods only or for a different gear ratio. At slow engine speed the relative supercharger -speed will be higher, giving' increased manifold pressure and thus some degree of responsiveness at low speed. In order to effect this, the gear box will be provided with a gear ratio which will give, for example, a supercharger speed of 1.5 times that employed in Scheme (a), but the fluid coupling will slip with the engine at 3000 rpm to give the same 1:1 ratio of Scheme (a), followed by no slip at low speed to give the effective ratio of 15:l1 at such low speed. 4

These three cases can be covered with one set of nomenclature: 7g = built-in ratio of gear box between engine and supercharger without slip. S = slip ratio of fluid coupling = ygN-n/ygN. ye = effective gear ratio between engine and supercharger = n/N = yg(l-S) n = supercharger speed rpm N = engine speed rpm rgN = maximum compressor speed (no slip) Then the work of compression can be given by: Work of Compression 2teNt = wCpTl(-1/k-l) 33000 rlcm t = compressor torque in lbs ft Tl = inlet air temperature ~abs R =compression ratio of supercharger w = lbs of air per min Cp = specific heat of air = 0 241 Btu/lb/OF C = isentropic efficiency of compressor nm mechanical efficiency of compressor The work to be provided by the engine to the driving element of the coupling with slip thus becomes: Work from engine 2 Nt - x Work of compressor 33000 Ye 1 w CpTl(R-/ ) 7e t ctlm 5

It must be remembered that there is always some slip in a fluid coupling even if completely filled with fluid. This in general amounts to about 3%. It follows that for the fully engaged condition, when an effective gear ratio of say 3:1 is desired, the design ratio yg must be 3.09 in place of 3.0, and the slip ratio at maximum compressor speed becomes S = 0.03, which in turn affects the work to be supplied the compressor by the engine, even in fullspeed condition. B 0C I _______cp! c A i___ Volume Flow cu ft/min Fig. 2. Typical performance map of a displacement supercharger with compression. The displacement compressor has a small variation of mass flow as the delivery pressure varies at constant speed, as can be seen from Fig. 2. If it is assumed that the volumetric efficiency and leakage due to pressure are constant, then the delivery would also be constant for any given speed and, to a first approximation (for relatively moderate engine-speed changes), the to the compressor rpm. Accepting these rather limiting assumptions initially, an approximation to the three schemes can be made with a moderate amount of calculation, and one can then judge if a mor e comprehensive method is justified.

Consider an engine of displacement D cu ins/cyldr, having "n" cylinders operating at N rpm, with a compression ratio of 17:1, then the air required to fill the cylinder and scavenge the clearance space is given by Air flow = (D + ) x N x n cu ins/min 16 2 053 nDN cu ins/min = 0.0184 nDN cu ft/hr = 0.0184 nDN Pm lbs/hr, where pm = density of air in manifold Pm = pressure of air in manifold psia Tm = temperature of air in manifold ~abs. It follows that the volume flow at the supercharger inlet becomes Flow to supercharger = 0.0184 nDN P- 540o0 at P1 = 14.7 and T1 = 540~ 14-7 Tm = 0.676 nDN Pm cu ft/hr. Tm m With an aftercooler giving Tm = 200~F, the flow is simplified to: Flow to supercharger = 0.001025 nDNPm cu ft/hr at inlet = 0.01506 nDN(Pm/Po) cu ft/hro The lines of constant supercharger speed in Fig. 2 can then be replaced by straight lines, as shown by the dotted line AB. Using this approximation, one can then calculate the performance over varying conditionso Assuming that the supercharger flow is also proportional to the rpm, the performance diagram of the compressor then becomes that shown in Fig. 35 plotted on a percentage basis, where A.B represents the volume flow of the compressor at some constant speed. One can now superimpose on Fig. 3 the engine air flow requirements and, assuming constant volumetric efficiency for any speed (reasonable for highly supercharged engines) with aftercooling to a constant temperature, the engine flow also becomes a straight line. The engine flow lines for different per7

FO \ 3, // / \ 3.0'L 2.... 10 20 30 40 50 60 70 80 90 100 110 Percentage Air Flow at Inlet 1. i.~ G H A / - 0I 10 20 30 40 50 60 70 80 90 100 110 Percentage Air Flow at Inlet Fig. 3. Supercharger characteristics plotted on a percentage basis.

centages of full speed then become as shown in Fig. 3, the 100% engine flow intersecting the 100% supercharger flow line at the desired pressure ratio (2.6:1 in Fig. 3)With the assumptions made, the manifold pressure will now remain constant as engine speed is varied if the supercharger is directly geared to the crankshaft with a constant speed ratioo This is indicated in the diagram by the intersection of the 50% engine-speed line with the 50% supercharger speed line at Pm/Po = 2.6:1 as for full speed, similarly for other speeds. When using the variable fluid coupling, one can operate the engine at 100% speed while the supercharger is at 50% speed only, In this case, Figo 3 indicates that the manifold pressure ratio will drop from 2.6:1 to 1.55:1 and the work of the compressor will be reduced, as well as the engine effective mean effective pressure developedo The required question is: Has the fuel consumption of the engine improved for this load condition at 1.55:1 relative to the same load at 2,6:1, as it does with constant compressor speed? B, CALCULATED RESULTS A series of calculations was made for pressure ratios of 1.0 to 2.6:1 for each of three engine speeds-3000, 2300, and 1600 rpm-at a constant fuel air ratio of 0.047 (representing full throttle conditions at each speed), and with a constant air temperature of 200~Fo The variation in pressure ratio was achieved by variable slip in the fluid coupling drive of the chargero The results are recorded in Table I and Fig. 4, The required slip in the coupling is determined from Fig, 3, e.g., for 100% engine speed (3000 rpm) and a pressure ratio of 2o0:1, the charger must operate at 72,5% of its 100% speed, because at these two conditions the air flow to the engine is equal to the air delivered by the compressoro In order to represent completely the operating conditions over the complete range of HP required by the vehicle, the data of Table I and Fig. 4 must be extended to a lower HP range. The low pressure ratio of lo0:1 of Table I represents the condition of the supercharger slipping to such an extent that it is merely "windmilling" to pass the requisite amount of air to the engine with no increase in pressure. Further reduction in HP can still be achieved by reducing the fuel flow below the F/A ratio of 0.047 by closing the fuel pump throttle. This part-throttle performance with slip is shown in Table II. In order to compare these data with normal engine operation employing a fixed-ratio drive to the charger (Scheme (a)), the data of Table III were calculated. In this case, the operating line of the charger was that at constant pressure ratio of 2.6:1, since at 60% engine speed, for example, the 9

TABLE I FULL-THROTTLE PERFORMANCE WITH SLIP Manifold Temperature = 200'F (F/A = 0.047) Pressure Ratio 1.0 1.2 1.4 1.6 1.8 2.0 2.2 2.4 2.6 IMEP (gross) psi 117 142 167 194 219 244 272 294 324 Compressor Work psi 3.3 5.0 10.3 14.7 17.2 25.2 32.9 41.5 52.0 Comp. rpm percent 23.5 33.0 42.5 52.0 62.0 72.5 80.5 91.0 100 Eff. Ratio (Ye) 0.235 0.33 0.425 0.52 0.62 0.725 0.805 0.91 0.97 Work With Slip psi 6.o 15.1 24.2 28.3 27.8 34.7 40.9 45.6 53.5 Friction and Cooling Loss psi 43.0 44.0 45.o 46.0 47.0 48.o 49.o 50.0 51.0 BMEP (net) psi 68.0 82.9 97.8 119.7 144.2 161.3 182.1 198.4 219.5 BHP 161.0 196.0 231.0 283.0 341.0 382.0 428.0 469.0 520.0 Specific BHP lb/sec 307.0 318.0 318.0 345.0 368.0 369.0 377.0 379.0 386.0 Specific IHP (net) lb/sec 500.0 485.o 464.o 478.0 488.0 479.0 481.0 475.0 476.0 Air Flow lb/sec 0.524 0.618 0.726 0.827 0.928 1.035 1.135 1.237 1.45 Fuel/IHP/hr 0.339 0.350 0.363 0.354 0.346 0.353 0.353 0.354 0.355 Fuel/BHP/hr 0.552 0.534 0.53 0.466 o.460 0.458 0.451 0.444 0.438 Comp. rpm percent 15.0 22.5 30.0 37.5 45.5 53.5 620 68.5 76.6 Eff. Ratio (ye) 0.196 0.294 0.392 0.49 0.594 o.698 0.81 0.89 0.94 rH^~ o Work with Slip psi 5.0 17.0 26.3 30.0 29.0 36.1 40.6 46.6 53.5 0 fo Friction and c Cooling Loss psi 29.5 30.1 30.9 31.6 32.3 33.0 33.6 34.3 35.0 BMEP (net) psi 82.5 94.9 109.8 132.4 157.7 174.9 197.8 213.1 235.5 BHP 150.0 172.5 199.5 240.0 286.0 318.0 359.0 388.0 428.0 Specific BHP lb/sec 375.0 361.0 359.0 380.0 402.0 401.0 411. 408.0 416.0 0 Specific IHP (net) lb/sec 509.0 476.0 460.o 471.0 484.o 477.0 481.0 474.0 478.0 Air Flow lb/sec 0.40 0.477 0.556 0.632 0.712 0.793 0.874 0.95 1.03 Fuel/IHP/hr 0.333 0.356 0.368 0.359 0.35 0.355 0.352 0.357. 354 Fuel/BHP/hr 0.451 o.469 0.472 0.446 0.422 0.423 0.412 0.416 0.407 Comp. rpm percent 8.5 12.0 18.0 23.5 29.5 36.5 42.0 47.7 53. Eff. Ratio (e) 0.16 0.225 0.338 o.44 0.553.684 0.787 0.894 0.97 o Work with Slip psi 4.0 22.2 30.5 33.4 31.1 36.8 41.8 46.4 5.5 fo Friction and Cooling Loss psi 18.0 18.6 19.3 19.9 20.5 21.1 21.8 22.3 23.0 BMEP (net) psi 95.0 101.2 117.2 140.7 167.4 186.1 208.4 225.3 247.5 BHP 120.0 128.0 148,0 178.0 212.0 235.0 263.0 285.0 313.0 ~ Specific BHP lb/sec 435.0 386.0 383.0 403.0 428.0 426.0 434.0 431.0 46. o Specific IHP (net) lb/sec 518.0 457.0 446.o 460.0 480.0 474. 479.0 474.0 476. Air Flow lb/sec 0.276 0.331 0.386 0.441 0.496 0.551 0.606 0.661 0.717 Fuel/IHP/hr 0.327 0.37 0.379. 368 0.353 0.354 0.353 0.357 0.355 Fuel/BHP/hr o.389 0.438 0.442 0.42 0.396 0.398 0.390 0.393 0.388

^ - ---- - 0055 ^ — - " ^' ^ x -- x F IA= -0.047 ^ 0.50 3000 RPM _10.45 ^ 2300 RPI 0)4 E~~~~~~~~~~~~~~~~~~~ 1600 RPPM 2300 /.3r 3.0 3ig. Lt.?erfOX/ance /jtV~ &LI at th~ttRPN- t t 2.6 RPM000 ^ /<600RPM.^ /x^RPM 9 2.2 y <- ^ L^^ —-^r —^ — ~^500 300 400 200 0 "*.full'ttle openance..witi slig at Fig-4 efrac

TABLE II PART-THROTTLE PERFORMANCE WITH SLIP F/A Ratio 0.047 0.04 0.035 0.03 0.025 0.02 IMEP (gross) 117.0 112.0 99..0 85.0 68.0 Compressor Work 6.o 6.o 6.o 6.o 6.o Friction and Cooling 43.0 42.5 42.0 41.5 41.0 o o BMEP (net) 68.o 63.5 51.0 37.5 21.0 Og BHP (net) 161.0 150.4 121.0 88.8 49.7 tr \ Specific BHP/lb 307.0 287.0 231.0 169.5 94.8 L a Specific IHP/lb 500.0 479.0 421.0 357.0 280.0 pI 6 Air Flow lb/sec 0.524 0.524 0.524 0.524 0.524 Fuel/IHP/hr 0.339 0.301 0.299 0.302 0.321 Fuel/BHP/hr 0.552 0.502 o.545 o.636 0.948 F/A Ratio 0.047 0.04 0.035 0.03 0.025 0.02 IMEP (gross) 117.0 112.0 99.0 85.o 68.o 42.0 Compressor Work 5.0 5.0 5.0 5.0 5.0 5.0 Friction and Cooling 29.5 29.0 28.5 28.0 27.5 27.0 ~ o BMEP (net) 82.5 78.0 65.5 52.0 35.5 10.0 BEP (net) 150.0 141.8 119.0 94.5 64.5 18.2 o, Specific BHP/lb 375.0 554.0 297.0 236.0 161.1 45.5 1t 0 Specific IHP/lb 509.0 486.o 426.0 363.0 286.0 168.5 O Air Flow lb/sec 0.40 o.4 o.4 o.4 o.4 o.4 Fuel/IHP/hr 0.333 0.296 0.296 0.298 0.315 0.427 Fuel/BHP/hr 0.451 o.406 0.425 0.459 o.56 1.58 F/A Ratio 0.047 0.04 0.035 0503 0.025 0.02 IMEP (gross) 117.0 112.0 99.0 85.o 68.0 42.0 Compressor Work 4.o 4.o 4.o 4.0 4.o 4.0 Friction and Cooling 18.0 17.7 17.3 17.0 16.7 16.5 0 o BMEP (net) 95.0 90.3 77.7 64.o 47.3 21.5 \o BHP (net) 120.0 114.0 98.2 80.8 59.7 27.2 Specific BHP/lb 435.0 413.0 355.0 293.0 216.0 98.5 P g Specific IHP/lb 518.0 489.0 434.o 371.0 295.0 174.0 p? & Air Flow lb/sec 0.276 0.276 0.276 0.276 0.276 0.276 Fuel/IHP/hr 0.327 0.295 0.29 0.291 0.307 0.414 Fuel/BHP/hr 0.389 0.353 0.355 0.368 0.415 0.732 12

TABLE III PART-THROTTLE PERFORMANCE, NO SLIP F/A Ratio 0.015 0.02 0.025 0.03 0.035 0.04 0.047 Pressure Ratio 2.6 IMEP (gross) psi 63 133 188 228 258 288 324 Compressor Work psi 52.0.0 2.0 0 5 52.0 52.0 52.0 Friction and 8 Cooling psi 46.0 47.0 48.o 49.0 50.0 51.0 51.0 EBMEP (net) psi 0 34 88 127 156 185 221 " BHP 81.0 209 301 370 438 524 Specific BHP hp/lb 60.2 156 224 275 526 390 Specific IHP hp/lb 143.5 241.0 311 363 416 479 Air Flow lb/sec 1.345 1.345 1.345 1.345 1.345 1.545 1.345 r Fuel/IHP/hr 0.502 0.374 0.347 0.347 0.347 0.354 _ Fuel/BHP/hr 1.19 0.578 0.481 0.458 0.443 0.436 Pressure Ratio 2.6 IMEP (gross) psi 63 133 188 228 258 288 324 Compressor Work psi 52.52.0 552.2.0 52.0 52.0 52.0 Friction and o Cooling psi 33.0 33.5 34.0 34.5 35 55.5 36.0 BMEP (net) psi 0 47.5 102 141.5 171 200.5 236 BHP 86 185 254 311 365 429 " Specific BHP hp/lb 83.5 179 249 302 354 416 Specific IHP hp/lb 142.5 239 310 364 417 479 ~. Air Flow lb/sec 1.03 1.03 1.03 1.03 1.03 1.03 1.03 o Fuel/IHP/hr 0.506 0.577 0.348 0.346 0.346 0.353 Fuel/BHP/hr 0.864 0.505 0.433 0.41 0.407 0.407 Pressure Ratio 2.6 IMEP (gross) psi 63 133 188 228 258 288 324 Compressor Work psi 52.0 52.0 52.0 52.0 52.0 52.0 52.0 Friction and 0 o Cooling psi 20.0 20.5 21.0 21.5 22.0 22.5 23.0 BMEP (net) psi 0 60.5 115.0 154.0 184.0 213.5 249.0'I BHP 0 76.5 145.5 195 233 270 315 Specific BHP hp/lb 106.8 203 272 325 377 439 Specific IHP hp/lb 143 240 0 1064 47 480 j Air Flow lb/sec 0.717 0.717 0.717 0.717 0.717 0.717 0.717 U Fuel/IHP/hr 0.503 0.375 0.348 0.346 0.346 0.353 H Fuel/BHP/hr 0.674 0.444 0.397 0.588 0.582 0.385 15

charger speed was also 60%, giving an intersecting point at 2.6:1 ratio from Fig. 3. The data of Table III record the results obtained for partload performance without slip; these data are also plotted on an F/A ratio base in Fig. 5. In order to simplify the evaluation of fixed and variable drive displacement superchargers, the data of Tables I, II, and III have been plotted on a BHP base in Fig. 6, where the solid lines record the fixed drive conditions and the dotted lines the variable drive conditions. The dotted lines show the data of Table I obtained by slipping the coupling until the pressure ratio reaches o10:1, then adding the data of Table II as the throttle is closed (AB represents slip at 3000 rpm, and BC throttle closing at 3000 rpm, similarly for the other speeds). Figure 6 reveals that the hypothetical 500-HP engine showed a slight loss in fuel economy, amounting to a maximum of 2.53 at 3000 rpm in the range of 350 to 500 HP; but from 350 HP down, there are savings in fuel for a given HP from zero at 350 BHP to 40% at 100 HP. At 2300 rpm there are savings of 42% at 75 HP reducing to zero at 225 HP, while at 1600 rpm there is a 45% reduction in fuel flow at about 60 BHP, reducing to zero at 160 HP. At all speeds there is a loss of 2-3% in the higher HP ranges. The problem now arises as to the average conditions under which operation occurs. Assuming that the power requirements under the conditions encountered in service do not exceed 50% to 60% of the maximum at any engine speed, the operation of the engine with a slipped compressor would show considerable improvement in fuel consumption relative to a similar compressor driven at a fixed ratio to the engine; and at the same time there is equal full-load HP available for all emergency conditions. C. RESPONSIVENESS By accepting the idea of a fluid coupling drive, one can then design into the compressor drive a two-speed arrangement to allow for normal power conditions, employing a low-speed supercharge drive with slip for economy, plus a high-speed drive for high power with responsiveness. Figure 7 is a diagram of one such a device. A compound wheel, D, driven by the engine crank can rotate two halves of the fluid couplings, A. and B, at two speeds in the ratio of say lo0 to 1.5. The other halves of the couplings are located back to back in casing C surrounding the couplings. Facilities will be provided (shown only for coupling A.) to supply oil to either coupling as desired. When coupling A is filled with oil, the high-speed gear ratio will be engaged and the supercharger will run at high speed with some slip in coupling Bo Because this coupling will be drained of oil, there will be only the small drag caused by the low oil flow required to lubricate the partsO By emptying A and filling B, the low speed is engaged for normal 14

Pm /Po = 2.6 Tm=2000 F 500 - 1.2 -- 1.1 400 - -- 1.0 |Ioo~~l % /1~ /-.0.9 x 300 - 0.8 0.7 Z \\ \ I l I I 200 0.6 Specific Fuel 0 0.5 3000 R PM 100 - 0.4 1 01600 R PM 0.3 0.01 0.02 0.03 0.04 0.05 F/A Ratio Fig. 5. Performance with fixed drive to supercharger. 15

1.0 1.0 \ | C*II~ \ ~Comparison of Operation 0.9 \ Fixed Ratio Charger Drive' V \\ --- Variable Ratio Charger Drive 0.8 0.7 I' A0.3 Fig an 0.2 J.. 3000 R. P. M. 0.3 0.2 0 100 200 300 400 500 B. H. P. Fig. 6. Comparison of fixed and slipping drive performance.

C High Speed Drive To'nBi:^ SLor Low Speed Drive To Supercharger H r D Oil Supply Driven Off Crankshaft Fig. 7. Supercharger drive for responsiveness.

operation. For the slip condition at low power, the coupling B can be drained to give the desired supercharger speed. This two-speed drive via a fluid coupling, then, would be superior to a low-speed drive with clutches, since the fuel savings indicated by Fig. 6 can still be achieved by slipping the coupling. One could employ just one coupling with the high-speed ratio, letting it slip sufficiently to give part-load performance. In this case, the work of compression would be high and the slip great, thereby increasing the fuel consumption to a value higher than that for the two couplings in one casing. Detailed analyses of these two methods would be necessary to make a final selection With the preceding rough indication of how to design such a drive, calculations were made to ascertain the engine capabilities with such a twospeed drive. These calculations are shown in Table IV, and the speed and flow conditions for the engine and charger are shown in Fig. 35 The action is as follows: The low-speed condition at full HP without slip is still represented by the intersection of the 100% engine and supercharger lines, and normal performance with slip at this gear drive is shown in Fig. 6 by the dotted lines. When the vehicle encounters extreme resistance, the high-speed coupling is filled as necessary to increase supercharger speed, and simultaneously the fuel supply is increased proportionally. The result is that as the engine speed is reduced due to increasing load, the supercharger ratio will increase along the line CD, reaching 355:1 at 1600 rpm of the engine, permitting a greatly increased torque development; this is in the case where the two drive gears are in the ratio of 15:lo If a 1.88:1 ratio were employed, the line CE would be followed, reaching a supercharger pressure ratio of 4.1:1l If desired, one could also apply the high-speed ratio at 3000 rpm, however the strain on the engine would be great, and the 1.88:1 speed ratio would still employ only the 100% supercharger speed. Thus, increased loads due to high speed on this unit would not be encountered. Admittedly, such increase in manifold pressure as would be achieved with the high gear ratio would impose additional gas loads on the engine. Using the VoCo.R principle of Continential Aviation and Engineering Corp. would avoid such increases. In Table IV, Columns 1 to 7 show the performance to be expected from 5000 to 1600 rpm along the increased pressure-ratio lines CD and CE of Fig. 5, and Columns 8 to 19 record performance along constant-ratio line DF; DH and FG indicate the portion where the fuel/air ratio is gradually reduced 18

TABLE IV RESPONSIVE PERFORMANCE WITH TWO-SPEED FLUID DRIVE DISPLACEMENT COMPRESSOR Pressure Ratio 2.6 2.6 3.0 3.5 3.2 4.1 4.1 IMEP (gross) 324 324 375 425 385 500 500 Compressor Work 52.0 61.6 74.6 67.2 88.0 Engine rpm 1200 3000 2300 1600 2300 1600 1200 Compressor rpm 1.0 1.0 1.2 1.5 1.31 1.88 1.88 Work with Slip 52.0 53.5 77.0 77.0 77-0 91.0 91.0 Friction and Cooling Losses 16.0 51.0 38.0 26.0 38.0 28.0 20.0 BMEP (net) 256 219.5 260 322 270 381 389 BHP (net) 242.5 520 472 407 490 481 368 Specific BEP 451.0 386 398 419 386 426 434 Specific IHP 478.0 476 456 453 440 457 456 Air Flow 0.538 1.345 1.185 0.97 1.27 1.13 0.848 Fuel/IHP/hr 0.354 0.355 0.371 0.374 0.385 0.37 0.371 Fuel/BHP/hr 0.375 0.438 0.426 0.404 0.439 0.398 0.39 F/A Ratio 0.047 0.047 0.047 0.047 0.047 0.047 0.047 Pressure Ratio 3.5 3.5 3.5 3.5 3.5 3.5 3.5 IMEP (gross) 385 350 330 - 250 185 425 Compressor Work 74.6 74.6 74.6 74.6 74.6 Engine rpm 1600 1600 1600 1600 1600 1200 Compressor rpm 1.5 1.5 1.5 1.5 1.5 1.5 Work with Slip 77.0 77.0 77.0 77.0 77.0 77.0 Friction and Cooling Losses 25.5 25.3 25.0 24.7 24.5 18.5 BMEP (net) 282.5 247.7 228 148.3 83.5 329.5 BHP (net) 357 313 288 187 105.5 312 Specific BHP 368 323 297 193 109 429 Specific IHP 401 356 329 225 141 453 Air Flow 0.97 0.97 0.97 0.97 0.97 0.728 Fuel/IHP/hr 0.359 0.354 0.329 0.40 0.51 0.373 Fuel/BHP/hr 0.392 0.391 0.365 0.367 0.66 0.394 F/A Ratio 0.04 0.035 0.030 0.025 0.02 0.047 Pressure Ratio 3.5 3.5 3.5 3.5 3.5 3.5 IMEP (gross) 385 350 330 250 185 147 Compressor Work Engine rpm 1200 1200 1200 1200 1200 1200 Compressor rpm 1.5 1.5 1.5 1.5 1.5 1.5 Work with Slip 77.0 77.0 77.0 77.0 77.0 77.0 Friction and Cooling Losses 18.0 17.5 17.0 16.7 16.5 16.0 BMEP (net) 290 256 236 156.3 91.5 54.0 BHP (net) 275 243 224 148 86.8 51.2 Specific BHP 378 334 308 203 119 70.4 Specific IHP 401 357 331 225 140 91.2 Air Flow 0.728 0.728 0.728 0.728 0.728 0.728 Fuel/IHP/hr 0.359 0.353 0.326 0.40 0.512 o.69 Fuel/BHP/hr 0.381 0.378 0.35 0.444 0.603 0.894 F/A Ratio 0.04 0.035 0.03 0.025 0.02 0.0175 19

from the full-load value of 0.047 to 0.0175, at which point the HP has been reduced to about 50 at both 1600 and 1200 rpm-the part-throttle performance. Figure 8 plots this responsive operation at full throttle, where HP, torque, and fuel consumption are shown for the 0.047 F/A ratio relative to the condition with zero-percent slip investigated in Table I. The original data of Table I indicate a gradually increasing torque from 1.0 to 1.18; with ye = 1.5, the ratio changes from 1.0 to 1.48; and with 7e = 1.88, it changes from 1.0 to 1o8.-a 78% increase in torque at 1200 rpm relative to that at 3000 rpm. This is not a high degree of responsiveness, but an examination of the outputs shows that the HP change at 1200 rpm is from 240 to 380, a fairly substantial increase at such a low speed. The fuel consumption data show a small increase in specific fuel consumption as 7e changes from 1.0 to 1.88. At 3000 rpm, all three conditions are the same because the blower speed also is constant; the relative speed ratio Ye does not reach 1.5 or 1.88 until the engine speed has been reduced to 1600 rpm. The maximum increase in specific fuel is 8.5% approximately and occurs at about 2300 rpm. When maximum responsiveness occurs, at 1600 rpm, the increase in specific fuel flow is only about 6% for a torque increase of 357 to 51-% Figure 9 shows the part-throttle performance for these two responsive conditions at 1600 and 1200 rpm; DH represents operation at 1600 rpm, and FG operation at 1200 rpm (from Fig. 3). The dotted fuel-consumption line in Fig. 9 represents performance at full throttle and 1600 and 1200 rpm, when ye = 188, giving 4.1:1 ratio, indicating the shift toward higher HP of the two part-throttle lines with little or no change in SFC. The degree of responsiveness is shown by the fact that 520 HP is developed at 3000 rpm and 480 HP at 1600 RPM if the 1.88:1 value of 7e is achieved. The problems of developing a supercharger, of the type being considered, to produce a pressure ratio of 4.1:1 would be great, though indications are that it would not be impossible. Its characteristics, as represented in Fig. 3, show that large speed increases are not necessary for a given size blower to achieve high pressure ratio, in contrast to the centrifugal machine. As already stated, the 4.1:1 ratio can be attained if, by means of the coupling, blower speed is kept at 100% while the engine speed is reduced to 53.3% of full speed, as shown by point E of Fig. 35 To demonstrate the relative performances, we have added to Fig. 9 one line representing a 3000-rpm performance with fixed drive ratio, employing the charger under discussion, and another line representing a fairly typical turbocharger performance. The former shows high specific fuel consumption at low loads due to the combination of high frictional losses plus constant compressor work, while the turbocharger suffers from low power output at low speed-approximately 90 BHP maximum at 1200 rpm for the centrifugal against 20

500 - FIA =0.047 e = 1.88 ox// 400 0 /X 300 -1.8 / ^71 ^ ^^ -1.88 ~ 20001 0.6.W 1.0 047 a -1e5 - 1 0.5..O" "0ye l.0o - 0.3 1200 1500 2000 2500 3000 RPM Fig. 8. Response operation at full throttle. 21

1.0 Pm I Po 3.5 and 4.1:1 0.9 0.8.<= \ \ Fixed Drive Reference Line 1-~ \ \ ~at 3,000 R PM 0.7 E 0.6 X <0 ~3.5:1.0 0.5 Ratio x~~~~~ 1200 RPM 1600 RPM mo - ^ 00 0.4 S -x^ ~* 1200RPM 4.1:1.0 Ratio 600RPM 0.3 Turbo-Charger Performance 100 200 300 400 500 B.H.P. Fig. 9. Part-load responsive performance.

210 HP at the same engine speed for Ye = 1.0, 240 with Ye = 1.5, and 285 HP when Ye= 1.88. The vehicle's performance in regard to acceleration, power to overcome obstacles, etc., would be greatly improved with the displacement blower at the expense of an increased fuel consumption in the normal power range up to about 200 to 250 BHP. D. CONCLUSIONS In normal operation on relatively level surfaces, a compression-ignition type engine fitted with a displacement supercharger with compression driven off the engine by means of a variable fill fluid coupling to permit speedratio variation possesses distinct advantages, as far as fuel flow is concerned, compared to the fixed drive ratio machine. The fuel flow up to about 250 HP, of a total of 500 HP, is superior by a relatively large margin, while in the 250- to 500-HP load range the loss is only a matter of some 2% relative to the fixed-ratio machine. It follows that operating conditions would play a great part in determining the type of supercharger to be used in a vehicle. Low-speed operation in difficult terrain, where high torque is necessary at low vehicle speed, favors the slip-speed blower, because this unit can be operated at the lowest speed at which it can develop the required HP, but can be driven at overspeed to give a good torque increase if conditions warrant; this overspeed can be obtained with a fairly simple two-speed device. When this type of supercharger is compared with the turbocharger, the two main differences are specific fuel consumption and power output or torque. The turbocharger engine has low torque at low speed, and thus low HP but good fuel economy, accompanied by some lack of acceleration and smoke when opening up from idle condition. At full power and speed, its power and economy are as good or better than the other designs, and at present it is capable of operating at higher manifold pressures than all other superchargers. The positive displacement machine will give superior low-speed torque and HP characteristics at the expense of considerably increased fuel flow, which can be reduced considerably if the slipping drive is applied. At the same time some degree of responsiveness is achieved. 23

II. COMBINATIONS OF THE MECHANICALLY DRIVEN SUPERCHARGER WITH TURBOCHARGERS A. DISPLACEMENT AND TURBOCHARGERS COMBINED This section will deal with the effect of combining a mechanically driven, positive displacement charger with a turbocharger. This combination is used because it can function turbocharged with maximum economy for normal operating conditions, and then for emergency conditions bring in the mechanically driven charger via a fluid coupling, or similar device, with only slightly increased fuel consumption. Calculations were made by combining the methods of Ref. 1 for turboand mechanically driven chargers. We assumed that the positive displacement machine was designed for various ratios-1.2, 1.4, to a maximum of 2.0:1and that it was then combined with a turbo of ratios o10 to 3.0 to give mean effective pressures up to approximately 400 psi; the overall fuel consumption was determined as before. The data of Table V were obtained from these calculations; the data of Table VI are for a typical turbocharged engine a lone Figure 10 plots the data of Tables V and VI on a base of BMEP for a constant engine speed of 3000 rpm; thus the base also represents HP to some scale. The case of the combined displacement and turbochargers shows that for a given BMEP the specific fuel increases as the ratio of the displacement compressor increases. This would be expected from the relative efficiencies of the two units. Secondly, the maximum BMEP achieved increases with the overall manifold pressure, but the straight turbocharged engine develops the higher mean pressure for any given value of the overall manifold pressure ratio; this is accompanied by the lowest specific fuel consumption over the whole load range. B, CONC LUSIONS From the standpoint of either power output or fuel consumption, there seems little merit in combining the characteristics of the displacement compressor with those of the centrifugal compressor if the maximum manifold pressure is to be held at some constant value, i.e,, at full load and highspeed operation. 25

TABLE V PERFORMANCE AT 3000 RPM Displacement Overall IMEP, ue Pressure Pressure (net), si b/IHP/hr Ib/BP/hr Ratio, Rm Ratio psi 1.2 135.6 91.6 0.345 0.510 1.68 194.2 147.2 0.329 0.432 2.16 247.9 199.4 0.340 0.422 1.2:1 2.64 305.5 254.5 0.338 o.405 3.13 365.1 312.1 0.335 0.392 3.6 425.8 370.8 0.330 0.378 1.4 163.6 118.7 0.334 0.461 1.96 228.3 180.7 0.335 0.424 2.52 289.9 239.9 0.340 0.410 3.08 356.5 304.0 0.330 0.387 3.64 417.0 362.0 0.341 0.393 4.2 477.6 419.3 0.343 0.391 1.6 181.0 135.0 0.345 0.462 2.24 253.5 204.5 0.346 0.428 2.88 328.9 276.0 0.342 0.407 1.:1 53.20 363.7 310.7 0.344 o.403 3.52 398.5 343.5 0.346 o.400 3.84 438.3 382.3 0.350 0.402 4.16 475.0 417.5 0.346 0.390 1.8 200.9 153.9 0.349 0.457 2.52 282.3 228.0 0.348 0.431 3.24 366.8 308.0 0.344 0.410 8:1 53.6 406.6 346.2 0.346 o.406 3.96 450.3 384.5 0.342 0.402 4.32 492.1 425.3 0.343 0.396 4.65 527.8 466.o 0.344 0.388 2.0 218.8 170.8 0.356 0.456 2.8 308.3 256.7 0.355 0.426 3.6 406.6 510.346 o.400 2.0o:1 4.0 449.3 392.3 0.347 0.397 4.4 495.1 436.1 0.347 0.394 4.8 548.9 486.9 0.342 0.386 5.2 588.7 524.7 o.344 0.386 TABLE VI TURBOCHARGED ENGINE Manifold e ManiflIMEP Fuel, Fuel, Pressure (net) Blb/IHP/hr lb/BHP/hr Ratio 1.6 189 143 0.326 0.432 2.0 234 186 0.330 0.414 2.4 279 229 0.334 o.405 2.8 338 287 0.521 0.377 3.2 373 320 0.330 0.386 3.6 424 369 0.328 0.376 4.0 474 417 0.327 0.370 4.4 522 464 0.326 0.366 4.8 568 508 0.526. 362 26

0.5 Rm 1.2 Rm= Pressure Ratio of Mechanically E 0.4, /Rm: 1.6 Driven Compressor ~. xRm= = *,=1.8''10 22 0 CO 0.3 100 200 300 400 500 B. M.E.P. p.s.i. Fig. 10. Full-throttle performance at 3000 rpm of combined chargers relative to a t urboc harger alone. chargers relative to a turbocharger alone.

As shown in Fig. 9, at reduced speed and load, substantial fuel would be saved by using a turbo alone, providing the power capability of the engine so fitted is sufficient for the purpose, or gear ratios can provide torque requirements. Due to the action of the positive displacement unit, the combined type of chargers would be of value only for increased torque capability at low speedo To capitalize on this, a drive should be provided to disconnect the displacement unit when not required, otherwise a heavy fuel flow penalty will result at part-throttle operation. The type of drive required-fluid coupling, clutches, etc.-would add complications over and above the need for two superchargers. The result would be increased engine bulk and complexity, with compensating advantages for short periods of operation only. The simpler, and thus generally more reliable, solution then would be to select (1), a high-ratio displacement compressor which produced high isentropic efficiencies, driven via a fluid coupling to give both high torque multiplication at low speed, coupled with slip for economy at part load, or (2), a turbocharger with variable turbine nozzles which would improve both lowspeed torque and fuel economy over the whole range of operation. Admittedly, (2) would not provide as much torque multiplication at low speed as (1). C. OBSERVATIONS The positive displacement machine with variable drive via a fluid coupling is particularly interesting in view of the recently published article, Differentially Compounded, Supercharged Diesel Engined(Ref. 3). The next report will analyze the differential drive machine and compare it with the type discussed in Section I of this report. The two are quite similar as far as the objectives are concerned, and a detailed analysis of the partload capabilities (not given in Ref. 3) as well as the full-load capabilities of the two systems must be made in order to evaluate the overall operation characteristics. 28

REFERENCES 1L E. T. Vincent and N. Tokuda, Engine Performance with Directly Driven Superchargers, Report No. 05847-1-T, Office of Research Administration, University of Michigan, Ann Arbor, September 1965. 2. E. To Vincent, Flexible Versus Responsive Engines, Report No. 04612-5-F, Office of Research Administration, University of Michigan, Ann Arbor, August 1962. 3. J. G. Dawson, W. G. Haywood, and P. Glamman, "Differentially Driven Superchargers for Diesel Engines," Institute of Mechanical Engineers, ADP, June 1964. 29

UNIVERSITY OF MICHIGAN 13 1l 35t 24