Tek silindirli içten yanmalı bir motorda sıkıştırma ve yanma periyodlarının bilgisayarla modellenmesi

Abstract

Bu çalışmada Otto tipi içten yanmalı motorların yanma odası geometrisini ve yalnızca sıkıştırma ve yanma zamanları gözönüne alınarak, krank mili açısına bağlı olarak silindir basıncının ve sıcaklığının değişimi, değişik motor devirleri ve hava/yakıt oranları için incelendi. Bu inceleme yukarıda belirtilen parametreleri simüle eden bir bilgisayar programı yardımı ile gerçekleştirildi. Bilgisayar yardımı ile modelleme günümüzde hemen her alanda gerek duyulan bir yöntemdir ve tercih ediliş nedenlerinin başında pratik oluşu gelir ki, bu da incelenen sistemde parametreler arası ilişkilerin her değişik durum için ayrı bir düzenekle modellenmesi gereğini ortadan kaldırışıdır. Ekte sunulan bilgisayar programı her ne kadar bir tek yakıt ve silindir geometrisi için çalışıyorsa da, gerekli eklemeler (yakıt türü ile ilgili) ve data değişiklikleri yapılarak farklı yakıtlar ve silindir geometrileri için de hesaplamaları gerçekleştirebilir. Böyle bir durum da bizi, her defasında başka boyutta bir motor ile farklı devir ve farklı yakıtların denenmek durumunda olması pratik zorluğundan kurtarır. Modellemede, yakıt girişi ve yanmış gazları egzost yolu ile atılışı ihmal edilmiş ve tek bir silindirin içinde sıkıştırma ve yanma zamanlarında gerçekleşen termodinamik olaylar 360 lik bir krank mili turu için hesaplanmıştır. Silindir duvar sıcaklıkları sabit kabul edilerek enerji denklemi içinde yer alan ısı geçişi terimi, sadece zorlanmış ısı taşınımı ile olduğu kabul edilerek Uoschni tarafından önerilen ısı taşınım korelasyonu aracılığı ile hesaplanmıştır. Sisteme ısı girişi ise adyabatik alev sıcaklıklarını ve buna bağlı olarak alev hızlarını yanma başlangıcından sonuna kadar her adımda bulan bir alt program yardımıyla hesaplanabilmektedir. Ayrıca sıkıştırma ve yanma periyotlarının her adımında (bizim sistemimizde bütün çevrim anlamına gelir) silindir içindeki yanmış ve yanmamış gazların termodinamik karakteristikleri hesaplanır. Alev hızları aracılığı ile her adımda yanan yakıt miktarı hesaplandıktan sonra bu miktar yakıtın alt ısıl değeri ile çarpılarak enerji denkleminde ısı girişi terimi olarak yerini alır. Her adımda bu denklemde ısı kaybı, ısı girişi ve iş terimleri yazılarak sonraki adımda silindir içindeki basınç ve sıcaklığın neler olduğu bulunur. Elde edilen değerler de bir sonraki adım için ilk şartlar olurlar ve bu döngü bütün çevrim boyunca devam eder. Bu program yardımı ile verilen hava/yakıt oranı ve devir sayısı için optimum ateşleme zamanı da krank mili açısına bağlı olarak belli bir hata yüzdesi ile hesaplanabilir. -iv-

In SI(Spark Ignition) engines the air and fuel are usually mixed together in the intake system prior to entry to the engine cylinder, using a carburetor of fuel injection system. In automobile applications, the temperature af the air entering the intake system is controlled by mixing ambient air uiith air heated by contact with the exhaust manifold. The ratio of mass flow of air to mass flow of fuel must be held constant at about 15 to ensure reliable combustion. The carburetor meters an appropriate fuel flow for the engine air flow in the following manner. The air flow through the venturi (a converging-diverging nozzle) sets up a pressure difference between the venturi inlet and throat which is used to meter an appropriate amount of fuel from the float chamber, through a series of orifices, into the air flow at theventuri throat. Oust downstream of the venturi is a throttle valve or plate which controls the combined air and fuel flow, and thus the engine output. The intake flow is throttled to below atmospheric pressure by reducing the flow area when the power required (at any engine speed) is below the maximum which is obtained when the throttle is wide open. The intake manifold is usually heated to promote faster evaporation of the liquid fuel and obtain more uniform fuel distribution between cylinders £ 1 ~J. Fuel injection into the intake manifold or inlet port is an increasingly common alternative to a carburetor. With port injection, fuel is injected through individual injectors from a low-pressure supply system into each intake port. There are several different types of systems: mechanical injection using an injection pump driven by the engine; mechanical, driveless, continious injection; electronically controlled, driveless injection. In this system, the air flow rate is measured directly; the injection valves are actuated twice per cam shaft revolution by injection pulses whose duration is determined by the electronic control unit to provide the desired amount of fuel per cylinder per cycle. An alternative approach is to use a single fuel injector located above the throttle plate in the position normally occupied by the carburetor. This approach permits electronic control of the fuel flow at reduced cost. Crank angle is a useful independent variable because engine process occupy almost constant crank angle intervals over a wide range of engine speeds. To maintain high mixture flows at high engine speeds (and -v-hence high power outputs the inlet valve, which opens before TC(Top Center), closes substantially after BC(Bottom Center). During intake, the inducted fuel and air mix in the cylinder with the residual burnt gases remaining from the previous cycle. After the intake valve closes, the cylinder contents are compressed to above atmospheric pressure and temperature as the cylinder volume is reduced. Some heat transfer to the piston, cylinder head, and cylinder walls occurs but the effect on unburnt gas properties is modest. Between 1G and kO crank angle degrees before TC an electrical dis-charge across the spark plug starts the combustion process. `A distributor, a rotating switch driven off the camshaft, interrupts the current from the battery through the primary circuit of the ignition coil. The secondary winding of the ignition coil, connected to the spark plug, produces a high voltage across the plug electrodes as the magnetic field collapses. Traditionally, cam-operated breaker points have been used; in most automotive engines, the switching is now done electronically. Aturbulentf lame develops from the spark discharge, propagates across the mixture of air, fuel and residual in the cylinder, and extinguishes at the combustion chamber wall. The duration of this burning process varies with engine design and operation, but is typically kO to 60 crank angle degrees. As fuel- air mixture burns in the flame, the cylinder pressure rises above the level due to compression alone. The pressure trace obtained from a motored or nonfiring engline is called motored cylinder pressure. Ue should note that due to the differences in the flow pattern or mixture composition between cylinders, and within each cylinder cycle-by-cycle, the development of each combustion process differs somewhat. As a result, the shape of the pressure versus crank angle curve in each cylinder, and cycle-by-cycle, is not exactly the same. There is an optimum spark timing which, for a given mass of fuel and air inside the cylinder, gives maximum torque. In this study, it will be tried to figure out the near boundries of an optimum spark timing for a reliable combustion by evaluating the two independent variables, the relative air-fuel ratio and R.P.M. as well as cylinder characteristics (Bore, stroke, etc..) for a single cylinder SI engine with a computer pragramme which is idealized by observing only two cycles; compression and combus-tion at given different spark timing, R.P.M and relative air-fuel ratio. More advanced (earlier) timing or retarded (later) timing than this optimum gives lower output. Called maximum brake-torque (MBT) timing, this optimum timing is an empirical compromise between starting combustion too early in the compression stroke -VI-(when the work transfer is to the cylinder gases) and cam pleting cobustian too late in the expansion stroke (and so louering peak expansion stroke pressures.) Abaut tua-thirds of the uay through the expansion stroke, the exhaust valve starts to öpen. The cylinder pressure is greater than the exhaust manifold pressure and a bloudouın process occurs. The burnt gases flouj through the valve into the exhaust port and manifold until the cylinder pressure and exhaust pressure equilibrate. The duration of this process depends on the pressure level in the cylinder. The piston then displaces the burnt gases from the cylinder into the manifold during the exhaust stroke. The exhaust valve opens before the end of expansion stroke to ensure that the blaudaun process does not last too far into the exhaust stroke. The actual timing is a compromise uıhich balances reduced uork trnsfer to the piston before BC against reduced uork transfer to the cylinder contents after BC £ 3 ]. The exhaust valve remains öpen until just after TC; the intake opens just before TC. The valves are opened and closed slauly ta avaid naise and excessive cam uear. Ta ensure the valves are fully öpen uınen piston velocities are at their highest, the valve öpen periods are often overlap. If the intake is throttled to belou exhaust manifold pressure, then backflou of burnt gases into the intake manifold occurs uıhen the intake valve is first opened. To illustrate the different types of engines in comman use, let's give some examples of praductian spark- ignition engines. Small SI engines are used in many applications: in the home, in portable pouıer generation, as outboard motarboat engines, and in motor-cycles. These are often single cylinder engines. in the above applications, light ueight, small bulk, and lau cast in relatian ta the pauer generated are the bost important charac-teristics; fuel consumption, engine vibratian and engine durability are less important. A single-cylinder engine gives only öne poujer stroke per revolution (tuo strake cycle) ör tua revalutians (faur-strake cycle). Hence, the tarque pulses are uidely spaced, and engine vibratian and smoothness are significant problems.Multicylinder engines are invariably used in automotive practice. As rated power increases, the advantages of smaller cylinders in regard to size, weight, and improved engine balance and smoothness point toward increasing the number of cylinders per engine. An upper limit on cylinder size is dictated by dinamic considerations: the inertial forces that are created by accelerating and decelerating the reciprocat-ing masses of the piston and connecting rod would quickly limit the maximum speed of the engine. Thus the displaced volume is spread out amongst several smaller cylinders. The increased frequency of power strokes with a multicylinder engine produces much smoother torque characteristics. Multicylinder engines can also achieve much better state of balance than single cylinder engines. A force must be applied to the piston to accelerate it during the first half of its travel from bottom-center to top-center. The piston then exerts a force as it decelerates during the second part of the stroke. It is desirable to cancel these inertia forces through the choice of number and arrangement to achieve a primary balance. Ue should note that, however, the motion of the piston is more rapid during the upper half of its stroke than during the lower half of (a consequence of connecting rod and crank mechanism). The resulting inequality in piston acceleration and deceleration produces correspond-ing differences in inertia forces generated. Certain combina-tions cylinder number and arrangement will balance out these secondary inertia force effects. Four-cylinder in-line engines are the most common arrangements for automobile engines up to 2.5-liter displacement [ k ~/. The /l arrangement, with two banks of cylinders set at 9G or more acute angle to each other, provides a compact block and is used extensively for larger displacement engines. For instance six cylinder (V6) engines provide smoother operation with three torque pulses per revolution. The in-line arrangement in a long engine, however, giving rise to crankshaft torsional vibration and making even distribution of air and fuel to each cylinder more difficult. The V6 arrangement is much more compect, and provides primary balance of reciprocat ing components. liJith the V engine, however, a rocking moment is imposed on the crankshaft due to the secondary inertia forces, which results in the engine being less well balanced than in-line version. The V8 and V12 arrangements are also commonly used to provide compact, smooth, low vibration, larger displacement, spark ignition engines. -vixi-The two-stroke cycle spark-ignition engine is used far small-engine applications where low cost and weight/power ratio are important and when the use factor is low. Examples of such applications are outboard motorboat engines, motorcycles, and chain saws. All such engines are of the carburetor crankcase-compression type which is one of the simplest prime movers available. It has three moving parts per cylinder: the piston, connecting rod, and the crank. The prime advantage of the two-stroke cycle spark-ignition engine relative to the four-stroke cycle engine is its higher power per unit displaced volume due to twice the number of power strokes per crank revolution. This is offset by the lower fresh charge density achieved by the two stroke cycle gas - exchange process and the loss of fresh mixture which goes straight through the engine during scavenging. Also, oil consumption is higher in two-stroke cycle engines due to the need to add oil to the fuel to lubricate the piston ring and piston surfaces. As a summary of this study and to give the general approaches to engine characteristics and their effects on engine variables, let's write down the main relations among engine characteristic: [ 5 *]. a) The effect of increasing the compression ratio on efficiency at a constant equivalence ratio is similar to that demonstrated by the constant k-constant volume cycle analysis. b) As the equivalence ratio is decreased below unity (the fuel-air mixture is made progressively leaner than stoichiometric), the efficiency increases. This occurs because the burnt gas temperatures after combustion decrease, decreasing the burnt gas specific heats and thereby increasing the effective value of k over the expansion stroke. The efficiency increases because, for a given volume expansion ratio, the burnt gases expand through a larger temperature ratio prior to exhaust; therefore, per unit mass of fuel, the expansion stroke work is increased. c) As the equivalence ratio increases above unity (the mixture is made progressively richer than stoichiometric), the efficiency decreases because lack of sufficient air for complete oxidation of the fuel more than offsets the effects of decreasing burnt gas temperatures which decrease the specific heats of the mixture. -IX-d) The mean effective pressure is prapartional to the product af (indicated fuel conversion efficiency multiplied by fuel-air equivalence ratia). This exhibits a maximum betüjeen PHI = 1 and PHI = 1.1, slightly rich af stoichiometric. Far PHI less than the value carresponding ta this maximum, the decreasing fuel mass per unit displaced valume more than offsets the increasing fuel conversicm efficiency. Far PHI greater than this value, the decreasing fuel conversion efficiency (due ta decreasing cambustian efficiency) tnare than affsets the increasing fuel mass [ B ]. e) l/ariations in initial pressure, inlet temperature, residual gas fractian, and atmospheric maisture fractian have cnly a modest effect on the fuel conversian efficiency. The effects af variatians in these variables an imep are mare substantial, hoiıiever, because imep depends directly an the initial charge density.