Valve Timing Diagram Of 4 Stroke Engine Diesel Engine

Principles of valve train operation

Krzysztof Jan Siczek , in Tribological Processes in the Valve Train Systems with Lightweight Valves, 2016

Four-Stroke Cycle

In four-stroke cycle engines, both SI and CI, there are four strokes completing two rotations of the crankshaft. These are respectively the suction or charging, compression, power/work or expansion, and exhaust strokes. The important variable characterizing operational conditions in each engine is the brake mean effective pressure (b _MEP), which is the mean effective pressure calculated from measured brake torque. It is defined by Eq. (2.1):

(2.1) ${\mathrm{b}}_{\mathrm{MEP}}={\mathrm{i}}_{\mathrm{MEP}}–{\mathrm{f}}_{\mathrm{MEP}}–{\mathrm{p}}_{\mathrm{MEP}}$

where

i_MEP is the indicated mean effective pressure, which is the mean effective pressure calculated from in-cylinder pressure—the average in-cylinder pressure over the engine cycle (720° in a four-stroke and 360° in a two-stroke). Direct i_MEP measurement requires combustion pressure-sensing equipment.

p_MEP is the pumping mean effective pressure, which is the mean effective pressure calculated from work moving air in and out of the cylinder due to inlet throttling losses and residual gases in outlet.

f_MEP is the friction mean effective pressure, which is the theoretical mean effective pressure required to overcome engine friction. It can be thought of as mean effective pressure lost due to friction.

Mean effective pressure is correlated with the peak pressure of gas in engine cylinders; however, such dependency is highly nonlinear and can be obtained from the measurement for a narrow class of engines or estimated using simulation models, which are very complex.

The mean effective pressure and peak pressure affect the force loading the seat faces, the valves, and their inserts, which determines friction between seat faces and their wear rate.

According to Ref. [14], for naturally aspirated SI engines, the maximum b_MEP is within the range 850–1050   kPa, at speed at which maximum torque is obtained. At rated power, b_MEP values are 10–15% lower. For boosted SI engines, the maximum b_MEP falls within the range from 1.25 to 1.7   MPa. For four-stroke CI engines, the maximum b_MEP is within the range 700–900   kPa for the naturally aspirated and 1.4–1.8   MPa for the boosted, respectively.

Four-Stroke Cycle SI Engine

In the SI engine, ignition is induced by sparks generated by spark plugs, where the operation cycle is adjusted to the engine speed and load using mechanical or computer-controlled ignition systems. Such adjustment is directly related to TDC positions and thus indirectly to the valve timing. In the SI engine, fuel is mixed with air, broken up into a mist, and partially vaporized. The compression ratio varies from 4:1 to 8:1, and the air–fuel mixture ratio varies from 10:1 to 20:1.

The four strokes of a petrol engine sucking fuel–air mixture are shown in Fig. 2.2.

Figure 2.2. The four strokes of an SI engine.

Four-Stroke Cycle CI Engine

In the CI engine, ignition takes place due to the heat produced in the engine cylinder at the end of the compression stroke. The four strokes of a CI engine sucking pure air are shown in Fig. 2.3. The compression ratio varies from 14:1 to 22:1. The pressure at the end of the compression stroke ranges from 30 to 45   kg/cm². The temperature near the end of the compression stroke is 650–800°C.

Figure 2.3. The four strokes of a CI engine.

The typical timing diagram for a CI engine is shown in Fig. 2.4.

Figure 2.4. The timing diagram for a CI engine.

From Ref. [15].

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Internal Combustion (Gasoline and Diesel) Engines

Robert N. Brady , in Encyclopedia of Energy, 2004

5 Four-Stroke Cycle Gasoline Engine: Basic Operation

Four-stroke cycle gasoline and diesel engines predominate as the global prime mover of choice. In gasoline engines, the introduction of electronic controls and GDI systems, in conjunction with dual overhead camshafts, distributorless ignition systems, variable geometry turbochargers, intercoolers, four-valve cylinder head designs, three-way exhaust catalytic converters, and the like, has greatly improved not only their thermal efficiency (fuel economy) but also the control of exhaust gas emissions.

In the four-stroke cycle, the piston moves through four individual strokes (two up and two down). The two upward-moving piston strokes are the compression and exhaust strokes, whereas the downward-moving piston strokes are the intake and power strokes within each cylinder. The individual strokes of the piston shown in Fig. 4 include the following:

Figure 4. Schematic operation of a four-stroke cycle gasoline engine. Courtesy of Prentice Hall.

•

The intake stroke where the piston moves down the cylinder. The intake valve(s) is held open by a rotating camshaft to permit an air/fuel mixture to enter the cylinder. This stroke can typically range between 240 and 260° obtained by opening the intake valve BTDC and then closing the valve ABDC.

•

The compression stroke where the piston moves up the cylinder to pressurize the air/fuel mixture while the intake and exhaust valves are held closed by valve springs.

•

The power stroke where the air/fuel mixture is ignited by a high-tension spark from the spark plug. The length or duration of the power stroke is controlled by opening the exhaust valve(s) before bottom dead center (BBDC). Typical duration of the power stroke averages approximately 140° of crankshaft rotation. The number of engine cylinders will determine how many power strokes are delivered within the 720° of crankshaft rotation in the four-cycle engine. For example, in a four-cylinder engine (720° divided by four), one power stroke will be delivered for every 180° of crankshaft rotation, whereas in a six-cylinder engine (720° divided by 6), one power stroke will be delivered for every 120° of crankshaft rotation.

•

The exhaust stroke, which begins when the exhaust valve(s) is opened before the piston reaches bottom dead center (BDC), typically between 30 and 40° of crankshaft rotation BBDC. The exhaust valves may remain open until 20° of crankshaft rotation after top dead center (ATDC) for a total duration of between 230 and 240° of crankshaft rotation.

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Tribochemistry of Lubricating Oils

In Tribology and Interface Engineering Series, 2003

Passenger car.

Most passengers cars have four stroke cycle gasoline engines. Current recommendations for these cars are oils which meet requirements for API Service SJ and ILSAC GF-2. These oils provide good protection against low temperature deposits and corrosion, protect against wear, and provide excellent protection against oxidation, thickening, and high temperature deposits under the most severe conditions of high speed operation or towing, even at high ambient temperatures.

The viscosities usually recommended are SAE 20W-20 or 30 for temperatures down to about 18°C, and SAE 20W-20 or 10W for lower temperatures. For extreme low temperatures, SAE 5W, 5W-20, or 5W-30 oils may be recommended. Multiviscosity SAE 10W-30 or 5W-30 oils are often recommended for year-round use, and are now the primary recommendation of several U.S. engine manufacturers. European engine manufacturers often recommend higher viscosities, particularly multiviscosity oils where SAE 20W-40 and 20W-50 oils are frequently preferred.

The recommendations for passenger car diesel engines are generally similar to those for four cycle gasoline engines; that is, oils for API Service CD or SJ/CD. Some manufacturers permit the use of multiviscosity oils, while others prefer single viscosity types. The rotary engines used in passenger cars generally require SAE 10W-30 oils for API Service SD or SE quality, and for two cycle gasoline engines, the oil used is usually of either SAE 30 or 40 viscosity and is formulated specifically for this service.

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Mechanical Heat Engines

Aldo Vieira da Rosa , Juan Carlos Ordóñez , in Fundamentals of Renewable Energy Processes (Fourth Edition), 2022

3.4 Four and Two Stroke Engines

In a reciprocating engine (Fig. 3.12), a piston moves back and forth in a cylinder. The linear motion of the piston is converted to the rotation of a shaft by means of a crankshaft mechanism. The piston moves between the top center (TC) and bottom center (BC). When the piston is at the top, the volume in the cylinder is minimum and it is called clearance volume, $V_c$ . The maximum volume, reached when the piston is at BC, is called total volume, $V_t$ , and the difference ( $V_t-V_c=V_s$ ) is the swept volume. Fig. 3.12 (right) illustrates a typical pressure-volume diagram for a reciprocating engine.

Figure 3.12. Geometry of a reciprocating engine (left). Pressure-volume diagram for a reciprocating engine (right).

Most reciprocating engines operate in a four stroke cycle. ^b In it, the piston sweeps the volume four times (up, down, up, down) while the shaft goes through two revolutions (see Fig. 3.13). The strokes are:

Figure 3.13. Four strokes: intake, compression, power (expansion), and exhaust (top). Pressure and volume plotted against crank angle (bottom).

•

Intake stroke: The cylinder moves from TC to BC. The intake valve is open and fresh fuel–air mixture goes into the cylinder.

•

Compression stroke: The valves are closed and the piston moves upwards, reducing the volume available and thus increasing the pressure. Near the end of this stroke, the combustion starts either triggered by a spark (in spark ignition engines) or self-induced by the pressure levels (as in compression ignition engines). Once the combustion starts, the pressure rises rapidly.

•

Power stroke (expansion): The piston descends from TC, pushing the crank mechanism and therefore rotating the shaft. The work done during this stroke is larger than the one required during the compression stroke. Once the piston approaches BC, the exhaust valve opens, allowing the combustion products to escape.

•

Exhaust stroke: The piston moves up towards TC, and with the exhaust valve opened the combustion products finish exiting the cylinder.

Fig. 3.13 (bottom) illustrates the pressure and volume variation as a function of crank angle. The volume follows a sinusoidal pattern, as expected in the slider-crank mechanism formed by the piston, connecting rod, and crank. The pressure exhibits a sudden rise during the combustion phase. Observe the timing in the intake valve opening and closing (labeled IVO, IVC) as well the exhaust valve opening and closing (EVO, EVC). You can imagine that timing of the valves and of the start of combustion play a major role in the engine operation.

A two stroke cycle is also possible (Fig. 3.14). In this case there are:

Figure 3.14. Schematic of a two stroke engine.

•

A compression stroke, in which the gases are compressed in the cylinder and fresh mixture comes into the crankcase. The combustion starts towards the end of this stroke when the cylinder approaches the top center.

•

A power stroke, in which the gases expand and the cylinder descends towards the bottom center. The gases are exhausted, and the fresh mixture that was in the crankcase is transferred towards the cylinder.

Two stroke engines have simpler valve systems and fewer moving parts, and are more power dense than four stroke engines. However, since they require oil to be mixed with the fuel, and some of the fuel mixture escapes unused, they pollute more than a comparable four stroke engine.

In order to simplify the analysis of gas cycles, it is common practice to use the so-called air-standard assumptions:

•

The working fluid is treated as air with ideal gas behavior. This avoids dealing with the chemistry of the fuel–air mixture and that of the combustion products.

•

The combustion process is replaced by heat addition, and the exhaust process by a heat rejection.

•

The cycles are closed with a fixed mass of air. There are no intake or exhaust valves.

•

All processes are treated as internally reversible.

In addition, if the specific heats are treated as constants and evaluated at ${\mathrm{25}}^{\circ }\mathrm{C}$ , the set of assumptions are called cold air-standard assumptions. We will, in some situations, extend the standard-air assumptions to gases other than air, by specifying the value of their specific heats.

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Engines

Hiroshi Yamagata , in The Science and Technology of Materials in Automotive Engines, 2005

1.1.2 The two-stroke engine

The two-stroke engine is similar to that of the four-stroke-cycle engine in its reciprocating mechanism. It uses the piston-crankshaft mechanism, but requires only one revolution of the crankshaft for a complete power-producing cycle. The two-stroke engine does not use inlet and exhaust valves. The gas exchange is implemented by scavenging and exhaust porthole openings in the bore wall. The upward and downward motion of the piston simultaneously opens and closes these portholes. The air-fuel mixture then goes in or out of the combustion chamber through the portholes. Combustion takes place at every rotation of the crankshaft.

In the two-stroke engine, the space in the crankcase works as a pre-compression chamber for each successive fuel charge. The fuel and lubricating oil are premixed and introduced into the crankcase, so that the crankcase cannot be used for storing the lubricating oil. When combustion occurs in the cylinder, the combustion pressure compresses the new gas in the crankcase for the next combustion. The burnt gas then exhausts while drawing in new gas. The lubricating oil mixed into the air-fuel mixture also burns.

Since the two-stroke engine does not use a valve system, its mechanism is very simple. The power output is fairly high because it achieves one power stroke per two revolutions of the crankshaft. However, although the power output is high, it is used only for small motorcycle engines and some large diesel applications. Since the new gas pushes out the burnt gas, the intake and exhaust gases are not clearly separated. As a result, fuel consumption is relatively high and cleaning of the exhaust gas by a catalytic converter is difficult.

In the past, petrol engines almost universally used ³ a carburetor. However, the requirements for improved fuel economy have led to an increasing use of fuel injection. In a petrol engine the fuel is normally injected into the inlet manifold behind the inlet valve. The atomized fuel mixes with air. When the inlet valve is opened, the combustible mixture is drawn into the cylinder. However, a recent development has occurred in direct injection petrol engines whereby fuel is injected directly into the combustion chamber, as with direct injection diesel engines.

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Theory and general principles

Malcolm Latarche , in Pounder's Marine Diesel Engines and Gas Turbines (Tenth Edition), 2021

The four-stroke cycle

Fig. 2.5 shows diagrammatically the sequence of events throughout the typical four-stroke cycle of two revolutions. It is usual to draw such diagrams starting at TDC (firing), but the explanation will start at TDC (scavenge). TDC is sometimes referred to as an inner dead centre (IDC).

Fig. 2.5. Four-stroke cycle.

Proceeding clockwise around the diagram, both inlet (or suction) and exhaust valves are initially open. (All modern four-stroke engines have poppet valves.) If the engine is naturally aspirated or is a small high-speed type with a centripetal turbocharger, the period of valve overlap (i.e. when both valves are open) will be short, and the exhaust valve will close some 10° ATDC.

Propulsion engines and the vast majority of auxiliary generator engines running at speeds below 1000   rev/min will almost certainly be turbocharged and will be designed to allow a generous throughflow of scavenge air at this point in order to control the turbine blade temperature (see also Chapter 10). In this case, the exhaust valve will remain open until exhaust valve closure (EVC) at 50–60° ATDC. As the piston descends to the outer or bottom dead centre (BDC) on the suction stroke, it will inhale a fresh charge of air. To maximize this, balancing the reduced opening as the valve seats against the slight ram or inertia effect of the incoming charge, the inlet (suction valve) will normally be held open until about 25–35° after BDC (ABDC) (145–155° BTDC). This event is called inlet valve closure (IVC).

The charge is then compressed by the rising piston until it has attained a temperature of some 550°C. At about 10–20° BTDC (before top dead centre) (firing), depending on the type and speed of the engine, the injector admits finely atomized fuel which ignites within 2–7° (depending on type again) and the fuel burns for 30–50° while the piston begins to descend on the expansion stroke, the piston movement usually helping to induce air movement to assist combustion.

At about 120–150° ATDC the exhaust valve opens (EVO), the timing is chosen to promote a very rapid blow-down of the cylinder gases to exhaust. This is done to (a) preserve as much energy as is practicable to drive the turbocharger and (b) reduce the cylinder pressure to a minimum by BDC to reduce pumping work on the 'exhaust' stroke. The rising piston expels the remaining exhaust gas and at about 70–80° BTDC, the inlet valve opens (IVO) so that the inertia of the outflowing gas, and the positive pressure difference, which usually exists across the cylinder by now, produces a throughflow of air to the exhaust to 'scavenge' the cylinder.

If the engine is naturally aspirated, the IVO is about 10° BTDC. The cycle now repeats.

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Motor Vehicle Emissions Control: Past Achievements, Future Prospects♦

John B. Heywood , in Handbook of Air Pollution From Internal Combustion Engines, 1998

1.6 DIRECT-INJECTION ENGINES, TWO-STROKES, AND DIESELS

Two other internal combustion engines are in widespread use besides the four-stroke cycle spark-ignition (SI) engine: the two-stroke gasoline engine and the diesel. In transportation, the two-stroke gasoline engine is used in the developing world for powering bicycles, small motor scooters, and motor bikes, because of its small size and weight, and low cost. The diesel dominates the truck engine market because its efficiency is substantially higher than that of the spark-ignition engine. In many countries the diesel has captured a significant share of the automobile engine market for similar reasons, especially in countries where fuel prices are high and where diesel fuel is taxed less than gasoline. While the fundamentals of the pollutant formation processes are similar in these other two engines, the details differ significantly and with the diesel there is a new problem—exhaust particulates.

The two-stroke cycle engine exhausts the burned gases from the cylinder largely by blowing in fresh air during approximately one-third of each crankshaft revolution as the crank moves through its bottom position. To make this scavenging process effective, a significant fraction of the fresh air flowing into the cylinder through the transfer ports in the bottom of the cylinder liner inevitably flows straight out of the exhaust ports (usually placed on the other side of the liner). With the simplest, small two-stroke SI engines that are carbureted, the gasoline is mixed with the air prior to entering the cylinder. So this short-circuiting of air directly through the cylinder results in a corresponding loss of fuel. This is a substantial fuel economy penalty (up to 25 percent), and results in very substantial hydrocarbon emissions. Thus, in cities with large numbers of motorized bicycles, motor scooters, motorcycles, and three-wheel taxis, the two-stroke cycle engine is an important source of emissions.

Substantial development efforts over the past 15 years have explored the potential of using direct in-cylinder gasoline injection to avoid this loss of fuel during scavenging. These efforts have been targeted at the automotive, the marine, and the motorcycle sectors. Figure 1.10 shows one of the more promising direct-injection technologies developed by the Orbital Engine Company applied to a crankcase-scavenged, two-stroke cycle engine. The necessary emission control with this concept is achieved by direct injection of gasoline into the cylinder, with an air-assist injector that achieves good dispersion of the fuel with very small drop sizes, after the rising piston closes the exhaust ports. Additional scavenging control is achieved with an exhaust flow control device (shown in the figure), a low-thermal-inertia, tuned, exhaust system, and a close-coupled oxidation catalyst to achieve fast light-off for HC and CO control. NO_x control is achieved within the cylinder. The two-stroke scavenging process leaves significantly more burned gases inside the cylinder, mixed with the fresh air, than does the four-stroke gas-exchange process. This additional residual burned gas in the in-cylinder fuel-air mixture reduces peak burned gas temperatures and the NO formation rate significantly.

Fig. 1.10. Features of the direct-injection two-stroke gasoline SI engine

(courtesy Orbital Engine Co.).

Whether or not this new direct-injection, two-stroke cycle technology will significantly penetrate the small engine/motorcycle market will depend on the cost of these fuel injection systems. Whether or not it becomes widely used in the automobile market will depend on the degree to which its durability and cost can be improved sufficiently to justify the development effort required to make this technology mass production feasible.

The diesel is the most efficient engine currently available and, consequently, is widely used in transportation (trucks, buses, railroads, and cars) when fuel economy is especially important. In the most efficient form of the diesel, the fuel is injected with a high-pressure injection system into a combustion chamber or bowl in the top of the piston toward the end of the compression process, as shown in Figure 1.11. The injected liquid fuel atomizes, forms a spray, vaporizes, mixes with the high-temperature air, and spontaneously ignites shortly after injection. Once combustion starts, it continues as additional fuel mixes with air to form a combustible mixture. Diesel emissions of hydrocarbons and carbon monoxide are low because combustion is almost complete and the engine always operates lean, with excess air. NO_x emissions are high, however, because burned gas temperatures are high. The three-way catalyst technology employed to good effect in the standard gasoline engine cannot be used to reduce NO_x levels in the diesel exhaust because the exhaust gas is lean rather than stoichiometric. Also, the fuel-air mixing process during combustion produces soot particles in the highly rich regions of each fuel spray. Some of this soot survives the combustion process unburned and absorbs high molecular weight hydrocarbons from the oil and fuel and sulfur as sulfate in the exhaust to form particulates.

Fig. 1.11. (a) Modern four-valve-per-cylinder turbocharged small high-speed direct-injection diesel (b) Fuel spray and bowl-in-piston combustion chamber characteristics of conventional (two valve, off axis inclined fuel nozzel, deep bowl, high swirl) and advanced (four valve, on-axis injector, shallow bowl, higher injection pressure-1600   bar, lower swirl) technology direct-injection diesel combustion systems

(courtesy of Ford Motor Co.), (courtesy of Mercedes-Benz AG).

Substantial control of NO_x emissions, and especially particulate emissions, from diesels has been achieved by modifications to the combustion process. Use of fuel injection equipment with very high liquid fuel injection pressures (~   2000   bar), and careful matching of the geometry of the bowl-in-piston combustion chamber, air motion, and spray geometry have significantly reduced soot formation by increasing fuel-air mixing rates. More careful control of lubricant behavior has reduced the high molecular weight hydrocarbon particulate component that is absorbed onto the soot. Use of low-sulfur fuels has reduced the sulfate component of the particulate. Oxidation catalysts in the diesel exhaust are increasingly being used to reduce further the soluble organic component of the particulate. NO_x reductions to date have been achieved by careful control of engine inlet air temperatures (e.g., turbocharged engines use an aftercooler to achieve low NO_x emissions), and with substantial injection retard to delay most of the combustion process to the early part of the expansion stroke. This latter strategy, of course, worsens fuel consumption by several percent.

While the diesel has made progress in reducing emissions (by about a factor of 3 to 4 for particulates, and a factor of 2 to 3 for NO_x), making this engine, which is the most efficient engine available, more environmentally friendly is an important task for engine developers and designers. Achieving substantially lower NO_x emissions is the major challenge. Some of this reduction could come from recycling exhaust gas, and a lesser amount from improvements in fuels. What is really needed is exhaust catalyst systems for NO_x reduction in the fuel-lean and low-temperature diesel exhaust environment. Lower levels of particulates will also be required.

In addition to these two-stroke and high-speed diesel engines, the direct-injection four-stroke spark-ignition engine is a potentially attractive new technology. Already in production in Japan, ⁶ this gasoline direct-injection engine offers improved fuel economy and is, therefore, one way to reduce vehicle CO₂ emissions. However, this engine's exhaust emissions are no better than those of the standard spark-ignition engine, and since it usually operates lean at light load, it requires a new catalyst technology to reduce NO_x.

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Vapor and Gas Power Cycles

Robert T. Balmer , in Modern Engineering Thermodynamics, 2011

13.18 Otto Cycle

The Stirling and Ericsson external combustion gas power cycles were originally developed to combat the dangerous high-pressure boilers of the early steam engines. The Lenoir internal combustion engine was simpler, smaller, and used a more convenient fuel than either of these engines, but it had a very poor thermal efficiency. Brayton managed to increase the thermal efficiency of the internal combustion engine by providing a compression process before combustion using the two-piston Stirling and Ericsson technique with a separate combustion chamber. But the ultimate goal of commercial internal combustion engine development was to combine all the basic processes of intake, compression, combustion, expansion (power), and exhaust within a single piston-cylinder apparatus. This was finally achieved in 1876 by the German engineer Nikolaus August Otto (1832–1891). The basic elements of the ASC model of the Otto cycle are shown in Figure 13.48. It is composed of two isochoric processes and two isentropic processes.

Figure 13.48. The Otto air standard cycle.

After several years of experimentation, Otto finally built a successful internal combustion engine that allowed all the basic processes to occur within a single piston-cylinder arrangement. The thermodynamic cycle of Otto's engine required four piston strokes and two crankshaft revolutions to complete, but it ran smoothly, was relatively quiet, and was very reliable and efficient. Otto's engine was an immediate success, and by 1886, more than 30,000 had been sold. They became the first serious competitor to the steam engine in the small- and medium-size engine market.

Initially, Otto's engine used illuminating gas (methane) as its fuel, but by 1885, many Otto cycle engines were already being converted into liquid hydrocarbon (gasoline) burning engines. The development of the ingenious float-feed carburetor for vaporizing liquid fuel in 1892 by the German Wilhelm Maybach (1847–1929) heralded the dawn of the automobile era. The German engineer Karl Friedrich Benz (1844–1929) is generally credited with building the first practical automobile, using a low-speed Otto cycle engine running on liquid hydrocarbon fuel, in 1885. He used engine exhaust heat to vaporize the fuel before it was fed into the engine.

Who Invented the "Otto" Cycle?

Unknown to Nikolaus Otto, the four-stroke cycle IC engine had already been patented in the 1860s by the French engineer Alphonse Eugene Beau de Rochas (1815–1893). However, Rochas did not actually build and test the engine he patented. Since Otto was the first to actually construct and operate the engine, the cycle is named after him rather than Rochas.

In 1878, the Scottish engineer, Dugald Clerk (1854–1932) developed a two-stroke version of the Otto cycle, producing one crankshaft revolution per thermodynamic cycle (it was like the Lenoir engine but with preignition compression). In 1891, Clerk went on to develop the concept of IC engine supercharging. This increased the thermal efficiency of the engine by further compressing the induction charge before ignition.

Although Clerk's two-stroke engine was inherently less fuel efficient than Otto's four-stroke cycle engine, it gave a more uniform power output (which is important only for single- or dual-cylinder engines) and had almost double the power to weight ratio of the Otto engine. The two-stroke Otto cycle (it never became known as the Clerk cycle) engine became successful as a small, lightweight engine for boats, lawn mowers, saws, and so forth.

The thermal efficiency of the Otto cycle is given by

${({\eta }_T)}_{\text{Otto}}=\frac{{({\dot{W}}_{\text{out}})}_{\text{net}}}{{\dot{Q}}_{\text{H}}}=\frac{{\dot{Q}}_{\text{H}}-\left|{\dot{Q}}_{\text{L}}\right|}{{\dot{Q}}_{\text{H}}}=1-\frac{\left|{\dot{Q}}_{\text{L}}\right|}{{\dot{Q}}_{\text{H}}}$

where, from Figure 13.48, $\left|{\dot{Q}}_{\text{L}}\right|=\dot{m}(u_{2s}-u_3)\text{and}{\dot{Q}}_{\text{H}}=\dot{m}(u_1-u_{4s})$ .

Then, the thermal efficiency of the Otto hot ASC is

${({\eta }_T)}_{\begin{array}{l}\text{Otto}\hfill \\ \text{hot}ASC\hfill \end{array}}=1-\frac{u_{2s}-u_3}{u_1-u_{4s}}$

For the Otto hot ASC, Table C.16a or C.16b in Thermodynamic Tables to accompany Modern Engineering Thermodynamics are used to find values for the specific internal energies. Since the processes from 1 to 2s and from 3 to 4s are isentropic, we use the v_r columns in these tables to find

$\frac{v_3}{v_{4s}}=\frac{v_{r3}}{v_{r4}}=\frac{v_{2s}}{v_1}=\frac{v_{r2}}{v_{r1}}=\text{CR}$

where $\text{CR}=v_3/v_{4s}$ is the isentropic compression ratio. If the intake temperature and pressure (T ₃ and p ₃) are known, we can find u₃ and v_{r 3} from the table. Then, if we know the compression ratio (CR), we can find

$v_{r4}=\frac{v_{r3}}{\text{CR}}\text{and}{v}_{r2}=v_{r1}\times \text{CR}$

We can now find u _4s and T _4s from the tables. However, to find u ₁, T ₁, u _2s, and T _2s, we need to know more information about the system. Consequently, the heat of combustion (Q _H/m = ${\dot{Q}}_{\text{H}}\text{}/\text{}\dot{m}$ ), maximum pressure (p ₁), or maximum temperature (T ₁) in the cycle is usually given to complete the analysis.

For the Otto cold ASC,

$\left|{\dot{Q}}_{\text{L}}\right|=\dot{m}(u_{2s}-u_3)=\dot{m}{c}_v(T_{2s}-T_3)\text{and}{\dot{Q}}_{\text{H}}=\dot{m}(u_1-u_{4s})=\dot{m}{c}_v(T_1-T_{4s}).$

Then,

${({\eta }_T)}_{\begin{array}{l}\text{Otto}\hfill \\ \text{cold}\text{ASC}\hfill \end{array}}=1-\frac{T_{2s}-T_3}{T_1-T_{4s}}=1-\left(\frac{T_3}{T_{4s}}\right)\left(\frac{T_{2s}/T_3-1}{T_1/T_{4s}-1}\right)$

The process 1 to 2s and process 3 to 4s are isentropic, so

$\begin{array}{ll}{T}_1/T_{2s}\hfill & =T_{4s}/T_3={(v_1/v_{2s})}^{1-k}={(v_{4s}/v_3)}^{1-k}\hfill \\ \hfill & ={(p_1/p_{2s})}^{(k-1)/k}={(p_{4s}/p_3)}^{(k-1)/k}\hfill \end{array}$

Since $T_1/T_{4s}=T_{2s}/T_3,$

(13.30) ${({\eta }_T)}_{\begin{array}{l}\text{Otto}\hfill \\ \text{cold}\text{ASC}\hfill \end{array}}=1-T_3/T_{4s}=1-{\text{PR}}^{(1-k)/k}=1-{\text{CR}}^{1-k}$

where $\text{CR}=v_3/v_{4s}$ is the isentropic compression ratio and $\mathrm{P}\mathrm{R}=p_{4s}/p_3$ is the isentropic pressure ratio.

Since $T_3=T_L$ but $T_{4s}<T_1=T_H$ , the Otto cold ASC thermal efficiency is less than that of a Carnot cold ASC operating between the same temperature limits (T ₁ and T ₃). Because the Otto cycle requires a constant volume combustion process, it can be carried out effectively only within the confines of a piston-cylinder or other fixed volume apparatus by a nearly instantaneous rapid combustion process.

Example 13.14

The isentropic compression ratio of a new lawn mower Otto cycle gasoline engine is 8.00 to 1, and the inlet air temperature is T ₃ = 70.0°F at a pressure of p ₃ = 14.7 psia. Determine

a.

The air temperature at the end of the isentropic compression stroke T _4s .

b.

The pressure at the end of the isentropic compression stroke before ignition occurs p _4s .

c.

The Otto cold ASC thermal efficiency of this engine.

Solution

a.

The isentropic compression ratio for an Otto cycle engine is defined as

$\text{CR}=\frac{v_3}{v_{4s}}={\left(\frac{T_3}{T_{4s}}\right)}^{\frac{1}{1-k}}$

from which we have

$T_{4s}=\frac{T_3}{{\text{CR}}^{1-k}}=T_3\times {\text{CR}}^{k-1}=(70.0+459.67\text{R}){(8.00)}^{0.40}=1220\text{R}$

b.

For the Otto cycle, the isentropic pressure and compression ratios are related by PR = CR ^k , where $\text{PR}=p_{4s}\text{/}{p}_3$ and CR = v ₃/v _4s . Then,

$p_{4s}=p_3{\text{CR}}^k=(14.7\text{psia}){(8.00)}^{1.40}=270.\text{psia}$

c.

Equation (13.30) gives the Otto cold ASC thermal efficiency as

${({\eta }_T)}_{\begin{array}{l}\text{Otto}\\ \text{cold}\text{ASC}\end{array}}=1-\frac{T_3}{T_{4s}}=1-{\text{PR}}^{\frac{1-k}{k}}=1-{\text{CR}}^{1-k}=1-{(8.00)}^{1-1.40}=0.565=56.5\%$

Exercises

40.

If the lawn mower in Example 13.14 is left outside on a cold day when T ₃ is reduced from 70.0°F to 30.0°F, determine the new temperature at the end of the isentropic compression stroke. Assume all the other variables remain unchanged. Answer: T _4s = 1130 R.

41.

If the clearance volume on the lawn mower in Example 13.14 is decreased such that the compression ratio is increased from 8.00 to 8.50 to 1, determine the new pressure at the end of the isentropic compression stroke. Assume all the other variables remain unchanged. Answer: p _4s = 294.1 psia.

42.

If the maximum temperature in the cycle (T _4s ) is 2400 R, determine the Otto cycle hot ASC thermal efficiency of this engine. Assume all the other variables remain unchanged. Answer: (η_T )_{Otto hot ASC} = 52.8%.

The actual pressure–volume diagram from an engine operating on a gas or vapor power cycle is called an indicator diagram, ¹⁰ and the enclosed area is equal to the net reversible work produced inside the engine. The mean effective pressure (mep) of a reciprocating engine is the average net pressure acting on the piston during its displacement. The indicated (or reversible) work output ${(W_I)}_{\text{out}}$ of the piston is the net positive area enclosed by the indicator diagram, as shown in Figure 13.49, and is equal to the product of the mep and the piston displacement, ${\mathit{V̶}}_2-{\mathit{V̶}}_1=\raisebox{1ex}{$\text{π}$}\!\left/ \!\raisebox{-1ex}{$4$}\right.{(\text{Bore})}^2(\text{Stroke}),\text{or}$

(13.31) ${(W_I)}_{\text{out}}=\text{mep}({\mathit{V̶}}_2-{\mathit{V̶}}_1)$

Figure 13.49. Mean effective pressure (mep) and indicator diagram relation.

The indicated power output ${({\dot{W}}_I)}_{\text{out}}$ is the net (reversible) power developed inside all the combustion chambers of an engine containing n cylinders and is

(13.32) ${({\dot{W}}_I)}_{\text{out}}=\text{mep}(n)\left({\mathit{V̶}}_2-{\mathit{V̶}}_1\right)\left(N/C\right)$

where N is the rotational speed of the engine and C is the number of crankshaft revolutions per power stroke (C = 1 for a two-stroke cycle and C = 2 for a four-stroke cycle). The actual power output of the engine as measured by a dynamometer is called the brake power ${({\dot{W}}_B)}_{\text{out,}}$ and the difference between the indicated and brake power is known as the friction power (i.e., the power dissipated in the internal friction of the engine) ${\dot{W}}_F$ , or

${({\dot{W}}_I)}_{\text{out}}={({\dot{W}}_B)}_{\text{out}}+{\dot{W}}_F$

therefore, the engine's mechanical efficiency η_m is simply (see Table 13.2)

(13.33) ${\eta }_m=\frac{{\dot{W}}_{\text{actual}}}{{\dot{W}}_{\text{reversible}}}=\frac{{({\dot{W}}_B)}_{\text{out}}}{{({\dot{W}}_I)}_{\text{out}}}=1-\frac{{\dot{W}}_F}{{({\dot{W}}_I)}_{\text{out}}}$

From Eq. (13.31), we can write

$\begin{array}{c}\text{mep}={(W_I)}_{\text{out}}/({\mathit{V̶}}_2-{\mathit{V̶}}_1)=\left({(W_I)}_{\text{out}}/m_a\right)/v_2-v_1\\ =\left[{({\dot{W}}_I)}_{\text{out}}/{\dot{m}}_a\right]/\left(v_2-v_1\right)\end{array}$

where m_a and ${\dot{m}}_{\text{a}}$ are the mass of air in the cylinder and the cylinder's air mass flow rate, respectively. The ASC (i.e., reversible or indicated, see Table 13.2) thermal efficiency of any internal or external combustion engine can now be written as

${({\eta }_T)}_{\text{ASC}}=\frac{{({\dot{W}}_{\text{out}})}_{\text{reversible}}}{{\dot{Q}}_{\text{in}}}=\frac{{({\dot{W}}_1)}_{\text{out}}}{{\dot{Q}}_{\text{fuel}}}=\frac{{({\dot{W}}_1)}_{out}/{\dot{m}}_a}{{\dot{Q}}_{\text{fuel}}/{\dot{m}}_a}$

where ${\dot{Q}}_{\text{in}}={\dot{Q}}_{\text{fuel}}$ is the heating value of the fuel. Combining these equations gives

$\text{mep}=\frac{{({\eta }_T)}_{\text{ASC}}\left({\dot{Q}}_{\text{fuel}}/{\dot{m}}_a\right)}{v_2-v_1}=\frac{{\left({\eta }_T\right)}_{\text{ASC}}\left({\dot{Q}}_{\text{fuel}}/{\dot{m}}_{\text{fuel}}\right)}{\left(A/F\right)\left(v_2-v_1\right)}$

where $A/F={\dot{m}}_a/{\dot{m}}_{\text{fuel}}$ is the air–fuel ratio of the engine. Now,

$v_2-v_1=v_1\left(v_2/v_1-1\right)=R{T}_1\left(\text{CR}-1\right)/p_1$

so Eq. (13.32) becomes

(13.34) ${({\dot{W}}_1)}_{\text{out}}=\frac{{\left({\eta }_T\right)}_{\text{ASC}}{\left(\dot{Q}/\dot{m}\right)}_{\text{fuel}}\left(DN{p}_1/C\right)}{\left(A/F\right)\left(R{T}_1\right)\left(\text{CR}-1\right)}$

where $D=n({\mathit{V̶}}_2-{\mathit{V̶}}_1)=\raisebox{1ex}{$\pi $}\!\left/ \!\raisebox{-1ex}{$4$}\right.{(\text{Bore})}^2\times (\text{Stroke})\times (\text{Number}\text{of}\text{cylinders})$ is the total piston displacement of the engine. Equation (13.34) allows us to determine the horsepower output of an ideal frictionless internal combustion engine, and when actual dynamometer test data are available, Eq. (13.33) allows us to determine the engine's mechanical efficiency.

Example 13.15

A six-cylinder, four-stroke Otto cycle internal combustion engine has a total displacement of 260. in³ and a compression ratio of 9.00 to 1. It is fueled with gasoline having a specific heating value of 20.0 × 10³ Btu/lbm and is fuel injected with a mass-based air-fuel ratio of 16.0 to 1. During a dynamometer test, the intake pressure and temperature were found to be 8.00 psia and $60.0°\text{F}$ while the engine was producing 85.0 brake hp at 4000. rpm. For the Otto cold ASC with $k=1.40$ , determine the

a.

Cold ASC thermal efficiency of the engine.

b.

Maximum pressure and temperature of the cycle.

c.

Indicated power output of the engine.

d.

Mechanical efficiency of the engine.

e.

Actual thermal efficiency of the engine.

Solution

a.

From Eq. (13.30), using $k=1.40$ for the cold ASC,

${({\eta }_T)}_{\begin{array}{l}\text{Otto}\hfill \\ \text{cold}\text{ASC}\hfill \end{array}}=1-{\text{CR}}^{1-k}=1-{9.00}^{-0.40}=0.585=58.5\%$

b.

From Figure 13.48a ,

${\dot{Q}}_{\text{H}}={\dot{Q}}_{\text{fuel}}={(\dot{m}{c}_v)}_a(T_1-T_{4s})={\dot{m}}_{\text{fuel}}\left(A/F\right){(c_v)}_a(T_1-T_{4s})$

and

$T_1=T_{\max }=T_{4s}+\frac{{\left(\dot{Q}/\dot{m}\right)}_{\text{fuel}}}{{\left(A/F\right)}_{\mathrm{mass}}{\left(c_v\right)}_a}$

Since process 3 to 4s is isentropic, Eq. (7.38) gives

$T_{4s}=T_3{\text{CR}}^{k-1}=(60.0+459.67){(9.00)}^{0.40}=1250\text{R}$

Then,

$T_{\max }=\frac{20.0\times {10}^3\text{Btu}/\text{lbm}\text{fuel}}{(16.0\text{lbm}air/\text{lbm}\text{fuel})[0.172\text{Btu}/(\text{lbm}air·\text{R})]}+1250\text{R}=8520\text{R}$

Since process 4s to 1 is isochoric, the ideal gas equation of state gives

$p_{\max }=p_1=p_{4s}\left(T_1/T_{4s}\right)$

and, since the process 3 to 4s is isentropic,

$T_{4s}/T_3{\left(p_{4s}/p_3\right)}^{(k-1)/k}$

or

${p_{4s}=p_3(T_{4s}/T_3)}^{k/(k-1)}=(8.00\text{psia}){\left(\frac{1250\text{R}}{520\text{R}}\right)}^{1.40/0.40}=172\text{psia}$

then,

$p_{\max }=(172\text{psia})\left[(8520\text{R})/1250\text{R}\right]=1170\text{psia}$

c.

Equation (13.34) gives the indicated power as

$\begin{array}{l}|{\dot{W}}_I{|}_{\text{out}}=\frac{(0.585)(20.0\times {10}^3\text{Btu}/\text{lbm})(260.{\text{in}}^{\text{3}}/\text{rev})(4000.\text{rev}/\min )(1170\text{lbf}/{\text{in}}^{\text{2}})/2}{(16.0)[0.0685\text{Btu}/(\text{lbm}·\text{R})](8520\text{R})(9.00-1)(12\text{in}\text{}/\text{ft})(60\text{s}/\min )}\hfill \\ =(132,00\text{ft}\cdot \text{lbf}/\text{s})\left(\frac{1\text{hp}}{550\text{ft}·\text{lbf}/\text{s}}\right)=241\text{hp}\hfill \end{array}$

d.

Equation (13.33) gives the mechanical efficiency of the engine as

${\eta }_m=\frac{{({\dot{W}}_B)}_{\text{out}}}{{({\dot{W}}_I)}_{\text{out}}}=\frac{85.0\text{hp}}{241\text{hp}}=0.353=35.3\%$

e.

Finally, the actual thermal efficiency of the engine can be determined from Eqs. (7.5) and (13.33) as

$\begin{array}{c}{({\eta }_T)}_{\begin{array}{l}\text{Otto}\hfill \\ \text{actual}\hfill \end{array}}=\frac{{({\dot{W}}_B)}_{\text{out}}}{{\dot{Q}}_{\text{fuel}}}=\frac{({\eta }_m){({\dot{W}}_{\text{I}})}_{\text{out}}}{{\dot{Q}}_{\text{fuel}}}=({\eta }_m){({\eta }_T)}_{\begin{array}{l}\text{Otto}\hfill \\ \text{cold}\text{ASC}\hfill \end{array}}\\ =(0.353)(0.585)=0.207=20.7\%\end{array}$

Exercises

43.

If the Otto cycle engine discussed in Example 13.15 has its compression ratio increased to 10.0 to 1, what would be its new Otto cold ASC thermal efficiency? Assume all other variables remain unchanged. Answer: (η_T )_{Otto cold ASC} = 60.2%.

44.

Find p _max and T _max for the Otto cycle engine discussed in Example 13.15 when the compression ratio is decreased from 9.00 to 8.00 to 1. Assume all other variables remain unchanged. Answer: p _max = 1040 psia and T _max = 8460 R.

45.

Determine the indicated horsepower in Example 13.15 if the engine's displacement is increased from 260. in³ to 300. in³. Assume all other variables remain unchanged. Answer: ( ${\dot{W}}_I$ )_out = 280. hp.

46.

Determine the mechanical efficiency of the Otto cycle engine in Example 13.15 if the actual brake horsepower is 88.0 hp instead of 85.0 hp. Assume all other variables remain unchanged. Answer: η_m = 36.3%.

The previous example illustrates that the Otto cold ASC analysis generally predicts thermal efficiencies that are far in excess of the actual thermal efficiencies. Typical Otto cycle IC engines have actual operating thermal efficiencies in the range of 15−25%. The large difference between the cold ASC (which contains at least one isentropic process) thermal efficiency and the actual thermal efficiency is due to the influence of the second law of thermodynamics through the large number of thermal and mechanical irreversibilities inherent in this type of reciprocating piston-cylinder engine. To improve its actual thermal efficiency, the combustion heat losses and the number of moving parts in the engine must be reduced.

What is the Smallest Internal Combustion Engine?

The Cox Tee Dee .010 model airplane engine (Figure 13.50) has the smallest internal combustion engine ever put into production. This amazing little engine weighs just under an ounce and runs at 30,000 rpm. The fuel is 10–20% castor oil plus 20–30% nitromethane mixed with methanol. With a bore of 0.237 in (6.02 mm) and a stroke of 0.226 in (5.74 mm), it has a power output of about 5 W.

Figure 13.50. Cox Tee engine.

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The piston

Hiroshi Yamagata , in The Science and Technology of Materials in Automotive Engines, 2005

3.1.1 Function

Figure 3.1 shows a typical assembly of pistons, connecting rods and a crankshaft for a four-stroke-cycle engine. The piston receives combustion power first and then, the connecting rod transmits the power through the piston pin. Figure 3.2 shows pistons used for two-stroke engines. Figure 3.3 shows both inside and outside views of a four-stroke engine piston. Figure 3.4 illustrates an assembly comprising piston, piston ring, and piston pin.

3.1. Forged pistons made of a continuously cast bar, connecting rods and a crankshaft for an in-line four-cylinder engine. The recess for the engine valve 〈5 valves〉 raises the compression ratio. Forging accurately stamps the recess shapes.

3.2. Two-stroke engine pistons made of a powder metallurgical aluminum alloy. Two-stroke engines normally install two piston rings 〈Chapter 4〉. The long skirt is necessary for exchanging gas.

3.3. Four-stroke engine piston, 〈a〉 inside and 〈b〉 outside. The skirt thickness is reduced to as small as 1.5   mm for light weight. Three piston-ring grooves are observable at both ends of the upper end. The piston-pin boss for the piston pin is positioned at the center. The high contact pressure exerted by the piston pin is likely to cause abrasive wear of the boss.

3.4. Nomenclature of each portion. The head and top-land areas reach the highest temperature because they directly contact with combustion gas. The piston ring placed into the ring groove is a spring sealing gas. The piston pin is a hollow tube made of carburized steel. The pin inserted in the pin boss is held at both ends with snap rings in order not to jump out. The piston pin revolves during operation.

A piston can best be described as a moving plug within the cylinder bore. Figure 3.5 summarizes its various functions. Firstly, the piston together with the piston rings, form the combustion chamber with the cylinder head, sealing the combustion gas. Second, it transfers combustion pressure to the rotating crankshaft through the piston pin and connecting rod. Third, in the two-stroke engine particularly, the piston itself works as the valve, exchanging gas (Chapter 2). Pistons also need to be mass-produced at low expense.

3.5. Functions of a piston for high output power

When exposed to a high-temperature gas, the piston must reciprocate at high speed along the cylinder bore. Even a small engine can generate a high power output at an increased revolution number. A petrol engine piston of 56   mm diameter can typically shoulder a load of about 20 kN and moves at a velocity as high as 15   m/s. If the number of revolutions is assumed to be 15,000   rpm, the piston material reaches the repetition number of 10⁷ times in 11   hours. This number roughly equates to the fatigue-strength limit of a material. To generate high power output, the piston should be designed to be as light as possible whilst retaining durability. For example, the piston in Fig. 3.1 weighs 170   g whereas those in Fig. 3.2 are 150   g on average. Such lightweight designs increase the stress on the piston material.

In comparison with petrol engines, diesel engines generate power at relatively low rotation, although a higher cylinder pressure is required. Figure 3.6 shows a direct injection diesel piston with a combustion bowl. The increased demand on diesels today requires high performance at high efficiency with low emissions. Most of the direct injection diesel engines used in cars in the year 2003 reached specific power outputs of up to 40   kW/L. These engines use a piston with a centrally located combustion bowl. The second generation of high-pressure injection systems (common rail and unit injector) has generated an increase in cylinder pressure up to peak values of 18   MPa. This increased power has raised the piston head temperature to as much as 350   °C at the combustion bowl edge. Compared to petrol engine pistons, diesels require a piston with a much higher strength to withstand these high temperatures.

3.6. Aluminum piston for a direct injection diesel engine. The edge of the combustion bowl is fiber-reinforced. The top ring groove is reinforced with a Niresist ring carrier.

Light aluminum alloy has been the most widely used material. The first aluminum-alloy piston appeared at the beginning of the 20th century just after the invention of the electrolytic smelting technology of aluminum in 1886. At that time, internal combustion engines used cast iron pistons. A key issue was whether aluminum, with a melting temperature as low as 660   °C, could withstand hot combustion gas. Figure 3.7 describes key requirements and reinforcement methods for aluminum and iron pistons.

3.7. Structures and manufacturing processes of automotive pistons.

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POLLUTANT FORMATION AND CONTROL IN SPARK-IGNITION ENGINES

JOHN B. HEYWOOD , in Energy and Combustion Science, 1979

6 WANKEL ENGINE EMISSIONS

The Wankel engine ⁶⁰ , ⁶¹ operating cycle has the same sequence of processes as the reciprocating SI engine four-stroke cycle. However, as a consequence of the Wankel geometry, there are important differences between the Wankel and conventional engine combustion processes which effect the emissions characteristics. In current production Wankel engines, the engine HC emissions are higher, CO about the same, and NO _x lower than an equivalent piston engine. However, only limited data on Wankel engine emissions is available in the literature. Since engine CO emissions depend on fuel–air equivalence ratio analogously to a conventional piston engine, we will not discuss them further.

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