Design and Simulation of Two Stroke Engines

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Design and Simulation of Two-Stroke Engines mb is the out-of-balance mass of the piston, connecting rod, etc., the maximum value of such forces, F, is given by: F = mbLstrpm2 It can be seen that this is apparently a linear function of the engine stroke, Lst, and so the lowest possible stroke length is used to minimize the vibration felt by the operator. Thus borestroke ratios of up to 1.4 are quite common in the design of handheld power tools, even though the potential bmep attainment of the engine is compromised by the reduced port areas available for any given set of port timings. As an alleviating factor, the highest crankcase compression ratio will always be produced by the largest bore-stroke ratio, and the scavenging of a simple loop-scavenged engine is relatively unaffected by bore-stroke ratio, both of which are vital factors in the design of this type of motor. Where there are no overriding considerations of the type discussed above, the two-stroke engine designer should opt for a bore-stroke ratio between 1.0 and 1.2. However, for diesel engines, where combustion is totally important, bore-stroke ratios are typically optimized at 0.9 for high-speed engines and can be as low as 0.5 for low-speed marine engines. In the latter special case, the values of bore and stroke are gross by comparison with the chainsaw engine, as the bore is about 900 mm and the stroke is some 1800 mm! 6.2.2 The width criteria for the porting The designers of high-performance two-stroke engines tend to favor the use of single exhaust and inlet ports, if at all possible. Although a divided port for exhaust or in,let is seen in many designs, it is always a highly stressed feature and the lubrication of the hot, narrow exhaust bridge is never an easy proposition at high piston speeds. Furthermore, as the need for large ports with long timings is self-evident for a high specific output engine, from Fig. 6.16 and the required specific time area criterion, every mm of bore circumference around the bdc position is needed for ports and not port bridges, so that the port opening periods can be kept as short as possible. Exhaust ports The designer of the racing engine will opt for a single exhaust port whose width is 75% of the bore dimension, and insert large corner radii to compress the piston ring into its groove before it passes the timing edge on the compression stroke. The net physical effect of the data chosen is seen in Fig. 6.16. The top edge itself can also be radiused to enhance this effect. To further enhance the blowdown effect without compromising the port width criteria, extra ports are added in the blowdown region, as shown in Fig. 5.1(b). In the case of the chainsaw engine of Fig. 6.2, and other such motors of a more modest bmep and power requirement, where the designer is not under the same pressure for effective port width at the bdc position, the typical criterion for the effective exhaust port width is to employ 55% of the bore dimension. For many such engines, a long and reliable engine life is an important feature of the design, so the ports are made to ensure a smooth passage of the rings over the port timing edges. (See Chapter 8 for further comments on the design of this port for noise considerations [8.14].) 432
CURRENT INPUT DATA FOR RACING ENGINE BORE DIAMETER= 54 LINER THICKNESS= 4 MAXIMUM EXHAUST WIDTH= 41 MAIN PORT, UP e AT 0 TARGET MT1=26.0 TARGET MT2= 5.0 ANGLE AM1=55 ANGLE AM2=50 SIDE PORT, UPS" AT 0 MAIN-SIDE BAR, MBAR= 4 SIDE-SIDE BAR, SBAR= 25 41 OUTPUT DATA MAIN PORT PLAN WIDTH= 19.8 SIDE PORT PLAN WIDTH= 12.0 TOTAL EFFECTIVE PORT WIDTH= 63.6 TOTAL EFFECTIVE WIDTH/BORE RATIO=1.18 Chapter 6 - Empirical Assistance for the Designer Fig. 6.16 Scavenge port layout for a piston ported racing engine. Inlet ports The inlet port width of a piston ported racing engine is deduced in much the same manner, with the port width to bore criterion for a single port being about 65% of the bore dimension, even though the piston rings may not intrude into the port. As all ports have to fit within the bore circumference, it is clearly necessary to distribute that available for the exhaust, transfer and the inlet ports. Often the inlet port width is conveniently made to coincide with the carburetor flow diameter, so the designer allows that dimension to be used, permitting a straight gas flow path for two sides of the intake duct. For a racing engine design with a wide single inlet port, the corner radii of the port need to be very generous around the bdc period by the upper timing edge of that port if a piston ring is not to be trapped, assuming that the rings actually so intrude. As for the exhaust port [8.14], Chapter 8 includes further comment on noise generation from the induction system with respect to this port design. Transfer ports The transfer port width for calculation purposes can be decided relatively easily. The geometrical layout can be effected by using Prog.3.4, although the experienced designer will know that it is mechanically possible to pack in a total effective transfer port width of between 1.2 and 1.35 times the bore dimension in most loop-scavenged designs. The designer, experienced or not, will wish to determine the port width criteria in some detail, and for this tedious exercise Prog.3.4 is invaluable. Fig. 6.16 shows the result of a few moments of time spent at the computer screen for a piston ported racing engine of 54 mm bore and stroke, confirming that the total port width 433
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Design and Simulation of Two-Stroke
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A Second Mulled Toast When as a stu
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Acknowledgments As explained in the
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Table of Contents Nomenclature xv C
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Table of Contents 2.10.1 Flow at pi
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Table of Contents 3.5.3.1 The use o
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Table of Contents 6.1.3 The effect
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Nomenclature NAME SYMBOL UNIT (SI)
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speed of rotation speed of rotation
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attenuation or transmission loss wa
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Design and Simulation ofTwo-Stroke
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Design and Simulation ofTwo'Stroke
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Chapter 4 Combustion in Two-Stroke
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Chapter 4 - Combustion in Two-Strok
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a stratilpl o tions to ivior. If m
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CD 30 -, W 20 - LU CO iS _J LU OC 1
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181 also x3 = — = 0.905 x6 = 2.17
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to be cur .3.20) .3.21) .3.22) for
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3.23) com- :on is hhas has a /eng-
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stroke .3.29) mean me to 3land ssio
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a LU z: en 3 CO a LU z EC D m O ho
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CD 1.2 -i 1.0 - di „„ LU 0.8 cc
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CO —> 111" cc LU < LU _J LU CC £
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Combustion in compression-ignition
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stroke ari more often lseful lental
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lat the ssi ;tribu- ELEVATION VIEWS
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CYLINDER HEAD TYPE IS CENTRAL BORE,
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CO E 50 40 - O O _i Hi > I CO 30 Z)
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CURRENT INPUT DATA FOR 2-STROKE 'SP
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Cylin- PP £ Pa- tics of Paper Chap
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vfitro- ;titi >tants 1970. i(In- Ch
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£2 = Pi P2 I Pi J Chapter 4 - Comb
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CHAINSAW AT 9600 rpm. Chapter 4- Co
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LLI •z. O N -z. 0.21 -, 0.20 DC 0
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Page 402 and 403: Design and Simulation of Two-Stroke
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0.35 • 0.34 LU ° 0.33 - c 03 DC
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Chapter 7 • Reduction of Fuel Con
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§ 2000
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y, ,6 Chapter 7 - Reduction of Fuel
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Chapter 7- Reduction of Fuel Consum
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INLET FOR r AIR AND FUEL Chapter 7
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Chapter 8 Reduction of Noise Emissi
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Chapter 8 • Reduction of Noise Em
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Chapter 8 - Reduction of Noise Emis
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• • / . • Appendix Listing of
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Active radical (AR) combustion and
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initial conditions, assumptions reg
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in two-stroke engine (schematic dia
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exhaust timing control valve, effec
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I Exhaust emissions/exhaust gas ana
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Exhaust systems (continued) untuned
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gas constant/universal gas constant
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in crankcase heat transfer analysis
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Mercury Marine air-assisted fuel in
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I PISTON POSITION (computer program
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Pressure wave reflection (in pipes)
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local acoustic velocity, 60 local M
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Roots blower in turbocharged/superc
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Scavenging (continued) scavenging e
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Simulation, engine See Computer mod
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temperature vs. crankshaft angle (c
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work per thermodynamic (Otto) cycle
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Design and Simulation of Two Stroke Engines

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