diesel engine fundamentals 2

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DieselNet Technology Guide

www.DieselNet.com. Copyright Ecopoint Inc. Revision 2001.02

Diesel Engine FundamentalsAbstract: The diesel engine is a compression-ignition internal combustion heat engine which can be

operated in both the four- and two-stroke cycle. The combustion process can be theoretically

modeled by applying thermodynamic laws of mass and energy conservation to the processes in the

engine cylinder. Basic design and performance parameters in diesel engines include compression

ratio, swept volume, clearance volume, a number of scavenging characteristics in two-stroke

engines, power output, indicated power, mechanical efficiency, indicated mean effective pressure,

brake mean effective pressure, specific fuel consumption, and more.

Introduction

Heat Engines

Operation of Reciprocating Internal Combustion Diesel Engines

Assessment of Engine Performance

Additional Performance Parameters and Their Definition

IntroductionNew emission regulations and performance requirements imposed on modern diesel engines arestrong motivators for new technology. Traditional design methods and parameters are continuously

being upgraded and new ones adopted to meet the competitive market pressures. However, to those

who are involved in modern engine designs, a good knowledge of the many aspects that not onlycontrol the engine performance, but also its emissions profile is imperative. This knowledge mustevolve from simple design principles to sophisticated interactions between various engine systems aswell as its post-combustion exhaust emissions control devices.

In this paper a review of the basic engine components, design parameters, geometric properties, andperformance issues are presented. These parameters are discussed with the view to maximizeperformance and fuel economy and minimize exhaust pollutants. While many of the principlesdiscussed in this paper may have application to other powerplants, the main emphasis is on dieselengines.

Heat EnginesHeat engines can generally be considered energy conversion machines. In essence, chemical energysupplied generally in the form of fossil fuel to a heat engine is combusted by mixing with oxygen

provided the right temperature is available and produces heat that is eventually converted to usefulwork. Heat engines can be classified as:

a. External combustion engines, or

b. Internal combustion engines.

Examples of the external combustion engine include the Stirling engine where heat is added to the

working fluid at high temperature and rejected at low temperature (compare "The Case..."). A network is produced by the working fluid. Heat added to the working fluid can be generated from

practically any heat source, such as burning fossil fuels, wood, or any other organic material. The

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steam engine is another example of external combustion systems. Heat added from an external sourceelevates water temperature until it is converted into steam that provides pressure and eventually thenet work. Steam engine powered cars in the U.S. between 1900 and 1916, however, they all butdisappeared by 1924. Reasons for their demise were mainly the size and number of the majorcomponents required for their operation such as furnace, boiler, turbine, valving, as well as theircomplicated controls [Taylor 1985a].

In internal combustion engines, the same combustible mixture that produces heat is responsible forproducing useful work. Two basic categories of internal combustion engines can be classified asrotating and reciprocating engines (compare "The Case..."). A very well known example of therotating engines is the Wankel engine which is also known as the rotary engine. A clear example ofthe second category is the reciprocating engines which in itself can be divided into two types: thetwo-stroke and the four-stroke engine. In internal combustion engines, the working fluid is thecombustible mixture of air and fuel. This fluid is responsible for releasing heat energy that is stored inthe fuel and producing useful work. Worthwhile goals common to the design and development of allheat engines include: maximizing work (power output), minimizing energy consumption, andreducing pollutants that may be formed in the process of producing work. Figure 1 shows the maincomponents of a reciprocating engine. Intake and exhaust valves are omitted for simplicity, howeverit is worth noting that in some two-stroke engine designs inlet and exhaust ports are used rather than

valves which are commonly used in four-stroke engines.

Figure 1. Basic Components of a Reciprocating Engine

Both internal combustion reciprocating engine categories (two- and four-stroke) may be equippedwith either a spark-ignited combustion system, also known as SI, or a compression-ignited systemknown as CI. Spark-ignited engines use a combustion cycle known as the Otto cycle, whilecompression-ignited engines use the Diesel cycle. Spark-ignited systems are characterized byhomogeneous or mostly well mixed charge of fuel and air. In this medium combustion is initiated bya spark and the flame propagates along a front from the spark location to the opposite side of thecombustion chamber. Diesel engines achieve their high performance and excellent fuel economy bycompressing air to high pressures then injecting a small amount of fuel into this highly compressed

air. The charge of air and fuel in these engines is described as heterogeneous meaning that the fuelremains separate from the air for a finite time in the combustion chamber. Heat generated due to thecompression of air in the cylinder causes the small amount of highly atomized injected fuel toevaporate. Mixing with the hot surrounding air in the combustion chamber, the evaporated fuelreaches its auto-ignition temperature and burns thus releasing the energy that is stored in that fuel.The auto-ignition temperature of fuel depends on its chemistry. Unlike the SI system, combustion incompression-ignited engines occurs at many points where the A/F ratio and temperature can sustainthis process.

Operation of Reciprocating Internal Combustion DieselEnginesReciprocating internal combustion diesel engines can be classified as two-stroke or four-strokedesigns. In the following sections, the operation of each classification is described and some of the




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parameters affecting performance are detailed.

Two-Stroke Engines

Principle of Operation

Figure 2. Two-Stroke Combustion Cycle

By definition, two-stroke engines require two strokes to complete their combustion cycle. Figure 2gives the details of the two-stroke combustion cycle. With the transfer and exhaust ports open, airunder slight pressure in the crankcase flows into the cylinder. The rising piston eventually covers thetransfer ports, thus trapping the inducted air into the cylinder. Further upward motion toward top deadcenter compresses the air where fuel is injected at the appropriate timing. Heat absorbed from thesurrounding hot compressed air causes the fuel to evaporate and mix with the air. Once the auto-ignition temperature is reached, combustion begins and causes the working fluid (combustiblemixture) to expand thus applying pressure on the surface of the piston thus producing useful work atthe engine output shaft. Meanwhile, fresh air flows into the crankcase to be compressed by thedescending piston on its way to bottom dead center. While descending, the piston uncovers theexhaust port starting the scavenging of the cylinder and causing a slight increase in crankcase

pressure. This increase in crankcase pressure causes the fresh air induction into the cylinder throughthe transfer port and resumes the cycle once again.

The two-stroke engine represented by the schematic in Figure 2 is not the only design of a two-strokeengine. Several other designs exist with various mechanical arrangements, but their principle ofoperation is essentially the same. Some two-stroke engines have inlet and exhaust ports placed at oneend of their cylinders, as shown in Figure 3 [Taylor 1985a].

Figure 3. Schematics of Two-Stroke Engines: (a) Opposite Side Port Arrangement; (b)




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Same Side Port Arrangement

Scavenging in Two-Stroke Engines

The process of purging exhaust gases from a previous cycle and filling the cylinder with fresh air fora new cycle is referred to as scavenging. The main method for scavenging two-stroke engines is by

using the pressure of the inducted fresh air to purge or displace the burned gases from the previouscycle. Generally, the greater the oncoming air pressure, the more complete the scavenging process.Therefore, better scavenging in two-stroke engines is achieved, in part by raising the pressure of freshair being inducted into the cylinder. This process is accomplished by using various devices such as

blowers, compressors, or pumps [Heywood 1988].

Scavenging in two-stroke engines is performed mainly by one of three methods:

Cross-scavenging

Loop-scavenging

Uniflow-scavenging

Figure 4. Scavenging Methods in Two-Stroke Engines: (a) Cross-Scavenging; (b) Loop-Scavenging; (c) Uniflow-Scavenging

There are other variations to these three methods, but these remain the principal implementations ofscavenging in two-stroke engines. Figures 4 (a), (b), and (c) illustrate the differences between thethree methods [Heywood 1988].

The uniflow scavenging system may use inlet ports and exhaust valves, as shown in Figure 4 (c), orinlet and exhaust ports in opposed piston engines such as shown in Figure 5 [Taylor 1985a].

Figure 5. Opposed Piston Two-Stroke Engine Design




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As the exhaust ports open (EPO), a rapid drop in cylinder pressure is observed. Evacuation of theexhaust gas through the exhaust ports is known as the blowdown process. A blowdown angle isdefined as the angle between EPO and the angle at which cylinder and exhaust system pressuresequalize. The inlet ports open (IPO) late into the blowdown process and allow fresh air to enter intothe cylinder. This step can begin only when the fresh air charge pressure exceeds that of the cylinder.Meanwhile, the piston reaches its bottom dead center (BDC) position and reverses its direction thusmoving towards its top dead center (TDC). The point at which the piston reverses its direction is

shown as BC (bottom center) in Figure 6.

Figure 6. Indicator Diagram of a Two-Stroke Cycle

The next event in the cycle is the inlet port closing (IPC) which takes place as the piston movestowards its TDC and covering the inlet ports. Further piston movement towards its TDC causes theexhaust ports to close (EPC) thus trapping the fresh charge and any residual exhaust in the cylinder.Thereafter, a steady but fast pressure rise is observed in the cylinder. This pressure rise is alsoaccompanied by a proportional increase in temperature. The crank angle during which both inlet andexhaust ports remain open is referred to as scavenging angle.The description given to the scavenging process so far was somewhat ideal and comprised distinctsteps. However, the actual scavenging in two-stroke engines is far less than ideal. In fact, duringscavenging not only does the fresh charge exchange heat with the residual gases, it also mixes with itand changes its chemical composition in the process. The final chemical make up of the mixture atthe end of the scavenging process plays an important role in the combustion quality as well as its

resultant emissions. An inherent loss in two-stroke engines results when some fresh charge escapesthrough the exhaust ports during scavenging. This phenomenon is often referred to as short-circuitingwhich leads to lower volumetric efficiency.

A two-stroke is usually smaller in size than a four-stroke engine having the same power output andtends to have higher specific power (power output for a given engine displacement) than its four-stroke counterpart. Two-stroke engines are generally less fuel efficient than four-stroke engines. Themain reason for this relative fuel inefficiency in two-stroke engines is poor scavenging and relativelylow volumetric efficiency.

Basic Performance P arameters in Two-Stroke Engines

Having described the two-stroke combustion cycle, we can now define several parameters that wouldhelp assess engine performance.




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Compression Ratio

Compression ratio is generally defined as the ratio of total (maximum) to minimum cylinder volume.According to this general definition, compression ratio is the product of dividing the cylinder volumewhen the piston is at BDC by the cylinder volume when the piston is at TDC. This definitioncomplies with the traditional or conventional understanding that applies to reciprocating internal

combustion engines.

It is worth noting however that the traditional definition may not reflect the actual compression ratioin two-stroke combustion systems. In fact, compression is almost non-existent until the piston covers

both inlet and exhaust ports in two-stroke engines. Therefore, a more accurate definition ofcompression ratio in two-stroke engines should limit the total cylinder volume to that in effect at thetime of exhaust port closure when the working fluid is trapped within the cylinder. In spite of thistechnicality, the traditional definition of compression ratio is still preferred over other alternatives.Therefore, compression ratio can be defined as follows:

Cr= (Maximum Cylinder Volume)/(Minimum Cylinder Volume)

Compression ratio is an important parameter by which engine efficiency can be determined. Whilethe overall efficiency of the engine may depend on a number of other parameters such as mechanicalefficiency, volumetric efficiency, and pumping efficiency to name a few, compression ratio has adirect impact on thermal efficiency. Equation 2 (also reported in The Case...) shows the mathematicalrelationship between compression ratio and thermal efficiency.

th = [1 - 1/Cr] 100

where

th- thermal efficiency,

Cr

- compression ratio, and

- the ratio of specific heat at constant pressure to the specific heat at constant volume.

It follows then that the greater the Cr, the higher the thermal efficiency of an engine.

Swept Volume

For an engine cylinder having a bore Bcyl

and a piston stroke Sp

, the swept volume (Vs) is defined by

the product of the cross-section area of that cylinder (Bcyl2/4) and the piston stroke (S

p), as follows:

Vs = (Bcyl2/4) Sp

Clearance Volume

Clearance volume (Vc) is the volume above the piston when it is at its TDC position. The swept and

clearance volumes (both terms are illustrated in Figure 7) are used to define compression ratio.Equation 4 is used to determine the geometric compression ratio of an engine:

Cr= (Vc + Vs) / Vc

(1)

(2)

(3)

(4)




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Figure 7. The Relationship between Swept and Clearance Volumes

Delivery Ratio

The delivery ratio is the ratio of the mass of fresh air delivered during the scavenge duration and themass of air required to fill the swept volume at ambient conditions. The mass of air required to fill the

swept volume at ambient conditions serves as reference and is conventionally calculated on the basisof the prevailing ambient pressure and temperatures conditions.

Dr= (Actual Air Mass Delivered During Scavenge) / (Reference Air Mass to Fill

Vs)

The delivery ratio defines how well an engine is able to fill the cylinder with air from the prevailingambient conditions. Obviously the more air is introduced into the cylinder the more oxygen isavailable to the combustion process. Even though the majority of the air inducted into the cylindermay not engage into the combustion process, still the opposite action of starving a diesel engine fromair may have devastating consequences on combustion efficiency. The evidence of a diesel engine

starved of air is usually in the form of black smoke that is very visible.

Scavenge Ratio

The scavenge ratio (Sr) is the ratio of the mass of fresh air supplied during the scavenge duration and

the mass of air required to fill the total cylinder volume at ambient conditions in naturally-aspiratedengines. The mass of air required to fill the total cylinder volume at ambient conditions serves asreference and is conventionally calculated on the basis of the prevailing ambient pressure andtemperatures conditions. Scavenge ratio should not be confused with the delivery ratio, the difference

being that in the scavenge ratio the reference is total cylinder volume as opposed to just the sweptvolume in the case of the delivery ratio.

Sr= (Actual Air Mass Delivered During Scavenge) / (Reference Air Mass to Fill

Total Cylinder Volume)

Naturally, a successful scavenge is required before we can introduce the maximum amount of air inthe cylinder. Therefore, good scavenging is a prerequisite for a good delivery as well as goodscavenge ratios.

Scavenge Efficiency

The working fluid in the cylinder at the point of exhaust port closure may consist of trapped fresh air(m

fa), residual air from a previous cycle that had not engaged in the combustion process (m

ra), and

exhaust that had not been scavenged (mex). The scavenge efficiency (Se) is the ratio of the trappedfresh air to the total mass trapped in the cylinder.

Se = mfa / (mfa + mra + mex)

(5)

(6)

(7)




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Purity of Charge

It is important to differentiate between the term purity of charge (Pu) and scavenge efficiency. These

terms are often confused with each other. Charge purity is the ratio of the air trapped in the cylinderat the start of combustion, to the total mass of the working fluid. The air trapped in the cylinder priorto combustion consists of the charge of fresh air trapped (m

fa) plus any residual fresh air that had not

engaged in prior combustion (mra).

Pu

= mfa

/ (mfa

+ mra)

From a combustion efficiency stand point it is important to utilize as much of the fresh chargeinducted into the cylinder as possible. In other words we would hope to have an extremely small m

ra

if any. Ideally, the highest purity of charge is achieved when all of the charge is consumed during thecombustion process.

Trapping Efficiency

The trapping efficiency (Te) is defined as the ratio of the charge retained in the cylinder to the charge

supplied. This definition applies at the point of port closure at which the charge is trapped within thevolume of the cylinder. The higher the T

e, the higher the fresh charge trapped together with its

oxygen content being made available to the combustion to follow.

Te = (Mass of Fresh Air Port Closing) / (Mass of Air Supplied to Cylinder)

There may be other design parameters that could be defined having direct implications on engineperformance and emissions characteristics. However, the preceding are the more common and oftenused.

Scavenging Characteristics in Various Two-Stroke Designs

It may be useful at this point to apply some of the definitions to various scavenging methods andexamine how well these concepts may meet design objectives. Figure 8 shows scavenging behaviorin large two-stroke diesel engines for three scavenging configurations.

Figure 8. Relationship Between Purity of Charge And Delivery Ratio for VariousScavenging Configurations

(8)

(9)




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The line showing perfect displacement indicates that the fresh charge has displaced all of the exhaustproducts from the prior combustion cycle and the entire cylinder volume is being occupied by a freshcharge. This line also implies that there was no mixing between the fresh charge and the products ofcombustion that were scavenged. On the other end of the spectrum, complete mixing may take placewhen the scavenging process allows enough interaction between the fresh charge and the products ofcombustion. Obviously, one should aim for perfect displacement as the design objective.Examination of the results shown in Figure 8 indicates that uniflow scavenging is the most effective

followed by loop scavenging and then cross scavenging.

Operation of Four-Stroke Engines

It takes four strokes to complete the combustion cycle in four-stroke engines. Figure 9 is a schematicrepresentation of the four-stroke combustion cycle as applied to a diesel engine. In the first stroke, theintake stroke, the piston moves from its position at top-dead-center (TDC) toward the bottom-dead-center (BDC). During most of the intake stroke, filtered air is inducted into the cylinder. In the secondstroke, air that was inducted into the cylinder is compressed by the piston moving back to TDC fromits starting position at BDC. This second stroke is known as the compression stroke where air in thatcylinder heats up to a temperature usually above the auto-ignition temperature of the fuel which isinjected into the cylinder near TDC. As the fuel burns, heat energy is released raising the pressure

inside a greatly reduced volume near TDC. This energy release produces pressure that is applied tothe top surface of the piston thus pushing it back toward its BDC.

Figure 9. Four-Stroke Diesel Operation

This stroke is known as the expansion stroke since it is through that expansion that power (pressure)was imparted to the piston and caused it to move to BDC. The expansion stroke is also known as thepower stroke for obvious reasons. It is also referred to by some as the work stroke since theexpanding gases were producing work by applying their pressure to the top of piston. The last of thefour strokes is the exhaust stroke where combustion by-products are exhausted into the exhaustsystem for evacuation into the atmosphere. In general, today's four-stroke diesel engines are equippedwith devices that enhance air charging and allow injecting additional fuel in amounts proportional tothe additional air inducted to improve specific power output.




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Figure 10. Pressure-Volume Diagram for Four-Stroke Naturally-Aspirated Engines

Changes in pressure, volume, temperature, and mixture composition occur during the four strokes.The pressure-volume diagram is often used to describe changes in pressure and volume in thecylinder during a complete cycle. Figure 10 illustrates pressure and volume changes for a naturally-aspirated diesel engine.

In Figure 10, intake and exhaust valve events are marked by points 1 through 4, where Point No. 1 is

the point at which intake valve opens, Point No. 2 is intake valve closing, Point No. 3 is exhaustvalve opening, and Point No. 4 is exhaust valve closing. It is important to note that both intake andexhaust valves remain open during the time between Points 1 and 4 as well as its equivalent crankangle duration. This period is referred to as valve overlap and plays an important role in engine

performance and its emission characteristics.

The intake valve closing occurs a few degrees beyond BDC to improve cylinder filling and therefore,the volumetric efficiency of the engine. Effective and rapid compression of the working fluid (air)

begins after intake valve closing as the piston travels from BDC to TDC. In naturally aspiratedengines, pressure inside the cylinder during the intake stroke is below atmospheric pressure.Restrictions through the air intake filter, air inlet piping, intake manifold, intake port, and intakevalve contribute to pressure loss and help reduce cylinder pressure to below atmospheric. Shortly

following combustion, the expansion stroke begins and is marked by a number of chemical reactionsand heat transfer processes while the piston travels from TDC to BDC. At Point No. 3, the exhaustvalve opens thus allowing some of the combustion products to go through a blowdown process as aresult of the pressure differential between the cylinder and the exhaust system. The remainder of theexhaust gases are expelled from the cylinder by virtue of the piston motion from BDC to TDC duringthe exhaust stroke.




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Figure 11. Pressure-Crank Angle Diagram For A Four-Stroke Diesel Engine

ID - ignition delay; EVC - exhaust valve closing; IVC - intake valve closing; TDC - top

dead center; BDC - bottom dead center; EVO - exhaust valve open; IVO - intake valveopen

Another way to illustrate the four-stroke cycle is through the Pressure-Crank angle diagram shown inFigure 11. In addition to the details outlined in the description of the pressure-volume diagram, the

pressure-crank angle diagram (Figure 11) highlights the point at which fuel is injection (I) as well asignition delay. During this delay fuel injected into the cylinder evaporates using heat from theworking fluid that had been compressed. The result of the heat transfer from the compressed air to thefuel is a reduction in the rate of pressure rise that is illustrated in Figure 2.11. Following the start ofcombustion, the rate of pressure rise increases dramatically and the combustion pressure peaks a fewcrank angle degrees past TDC. Factors controlling the rate of pressure rise include: the ignition delay,fuel quality, and the rate of injection. In many designs, the engine noise, vibration, and harshnesscharacteristics is often tied to the rate of pressure rise in the cylinder.

Together with the rise in cylinder pressure, cylinder temperature also increases and reaches its peak.The maximum combustion temperature depends on several factors including: fuel rate, fuel injectiontiming, fuel quality especially its calorific value and cetane number, initial cylinder pressure at intakevalve closing, and charge temperature.

So far intake and exhaust valve motion has been treated generally in relation to the specifics of thepressure-volume or pressure-crank angle diagrams. Earlier, valve overlap was cited as a parameterhaving major influence on engine performance and its emission characteristics. A more detailed lookinto the opening and closing of these two valves reveals more insight into their effect on volumetricefficiency and total charge composition. Starting with the exhaust valve, its opening takes place at a

point near the end of the power stroke. The timing of EVO should not be too early lest useful work islost, yet it is beneficial to effect the valve opening while cylinder pressure is at a level capable of

clearing exhaust products through the exhaust valve area during the blowdown time. To ensurecomplete clearing of the exhaust gases through the exhaust valve opening it is usually kept open efew degrees past TDC. It is worth noting that both valves do not reach their fully open positioninstantaneously, but have a finite length of time during which they move from a fully closed to fullyopen position. To make the best use of both valves they must be in their fully open position at thetime when they are exposed to the maximum pressure differential causing the working fluid to flowacross them. To ensure this aspect of the cylinder filling and emptying processes, the intake andexhaust valves are usually opened before the start of the intake and exhaust strokes, respectively.They are also held open a few degrees past their respective strokes for the same purpose. As theexhaust valve is opened exhaust gases are expelled by their own kinetic energy, thus reducing thework required by the piston during the exhaust stroke to expel the remaining exhaust gases. Ensuringthe expulsion of all exhaust products from the cylinder facilitates the induction of a fresh and

generous charge to sustain combustion of the tiny quantity fuel injected into the cylinder. The intakevalve opens towards the end of the exhaust stroke while exhaust gases are exiting through the exhaustvalve at high velocity. Their high exit velocity creates a pressure drop in the cylinder that helps drawthe fresh charge and maximizes induction. If the intake valve opening is too early, some of the




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exhaust products may actually flow back through its opening into the intake port and then on to theintake manifold. This portion will be inducted once again into the cylinder during the intake strokeand contaminates the fresh charge with exhaust. The proper function of both valves may requiretiming their opening and closing in a way that may interfere with the piston position. In such casestwo solutions may be used: valve recess in the cylinder head as shown in Figure 12 or valve cutoutsin the piston crown as shown in Figure 13. Therefore, it is extremely important to carefully design thetiming of intake and exhaust valve opening, closing, lift, flow area as well as intake and exhaust valve

overlap.

Figure 12. Valve Without Recess (a) and with Recess (b)

Figure 13. Valve Cutouts in Piston Crown

There are two basic types of four-stroke diesel engines: the direct- and the indirect-injected engines.The design and operation of both types was described in The Case for the Diesel Engine paper,where basic differences in their construction were explained. Recent interest in energy conservationhas caused most if not all new four-stroke diesel engine designs to adopt the direct-injection conceptfor its superior fuel economy and its future promise of low emissions.

In spite of the promising future for direct-injected diesel engines, the indirect injected engine (IDI)clearly dominates the passenger car market. This dominance is due to its relatively quiet and smoothoperation when compared with older direct injected diesels. In addition, IDIs have higher powerdensity than DIs and are characterized by lower nitric oxide emissions than its DI counterpart. IDIengines are more popular around the world than they are in the United States. The main reason fortheir popularity is their superior fuel economy (15-20%) over their gasoline competitors. Their fueleconomy advantage over gasoline engines is more prevalent in urban driving conditions where thespark-ignited engine is normally heavily throttled. Modern, port-injected gasoline engines may havefuel evaporation difficulties during cold weather operation, and that is another area where IDI engines




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may be more superior. This IDI dominance over DI engine in the passenger car market has recentlybeen challenged. In the mid-1980s, Ford Motor introduced its first production light-duty high speedDI diesel engine (HSDI)to the European market. The new breed of HSDI engines are even more fuelfrugal than their IDI predecessors and have already demonstrated an average of between 10 and 15

percent lower fuel consumption than IDI engines. When compared to gasoline engines, the fueleconomy advantage may approach 30 to 40%. New engine designs and materials used in producingnew DI engines for passenger car applications are proving that HSDIs can be made to operate with

very acceptable noise, vibration, and harshness (NVH) characteristics [Foulkes 1995].

Although IDI engines have lower combustion noise, less costly fuel injection system, higher powerdensity, and in general lower emissions levels, HSDIs have better startability in cold weather, lessheat rejection to the water jacket, more EGR tolerance, greater durability, and lower overall carbondioxide emissions [Foulkes 1995].

Other Classifications of Engines

So far we have treated reciprocating engines as either spark- or compression-ignited engines. Wehave even subdivided compression-ignition engines into two categories. By so doing, we obviouslyrun the risk of over simplifying engine categorization. Therefore, to put this potentialmisunderstanding to rest we need to mention here that there are many other ways by which enginescan be categorized. Engines can be categorized by the following characteristics:

Mobility: Engines can serve both stationary as well as mobile applications. Stationaryengines range from very small power outputs perhaps below 2 horsepower such as Hatzsingle cylinder engines to as high as 21,000 horsepower such as the Harland and Wolffuniflow-scavenged two-stroke engine pictured in Figure 14 [Blair 1996]. Stationaryengines may be used as power generation either as standby or continuous applications.They are also used in driving machinery such as large compressor that specialize inmoving bulk.

Figure 14. Harland and Wolff Uniflow-Scavenged Two-Stroke 21,000Horsepower Engine

(Courtesy: Gordon P. Blair)

Fuel: So far we have mainly addressed engines that use either gasoline or diesel fuels.Yet, fuels are not necessarily limited to these two, but include others such ascompressed natural gas (CNG), liquid propane gas (LPG), methanol, ethanol, biodiesel,




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broad-cut, kerosene, dimethyl-ether, and a number of others. The less common fuelsmay be used in special applications that may be located near an abundant supply of thatfuel.

Application: Engines may be installed in vehicles serving off road or on highwayapplications. Off road engines may cover multiple applications ranging from earth-movers, back hoes, and other construction equipment to agricultural machinery such as

combines, tractors, mowers and garden tools. On highway engines may power heavytrucks as well as passenger cars and mopeds. Other applications that do not quite fit thedescription of on highway or off road may include aircrafts, marine, locomotive, and

portable units.

Configuration: Engines are designed to fit a variety of installations. For many years in-line engine designs (see Figure 15a) were most popular for vehicular applications. Yet,in the mid-1970's, front-wheel drives became popular to provide more room in the newdownsized car designs. Transverse engine installations became common practice andled to rash of conversions to vee engine configurations (see Figure 15b) from the in-linedesigns. The vee configuration compressed the length of the engine and allowed its

packaging under the hood of the passenger car. Other configurations include

horizontally opposed piston (see Figure 16), radial (see Figure 17a) for aircrafts, anddelta (see Figure 17b) [Power 1964].

Figure 15. Engine Configurations: (a) In-Line; (b) Vee

Figure 16. Horizontally Opposed Piston Engine




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Figure 17. Engine Configurations: (a) Radial; (b) Delta

Valve/Port Design: The position of intake and exhaust valves or ports can be used todifferentiate among various engine designs. For instance, overhead cam design facilitatethe actuation of unit injector fuel injection systems. Actuation of the unit injector isoften accomplished by inserting an additional cam between the intake and exhaust

cams. Figure 18 shows a variety of cam arrangements as applied in practical engineembodiments.

Figure 18. Various Cam Arrangements

OHV - overhead valve; OHC - overhead cam; DOHC - dual overhead cam

Induction: Induction systems come in two basic designs; naturally-aspirated orturbocharged. Hence, engines that do not have devices to boost their charge are referredto as naturally-aspirated and those that have charge boost devices, such as turbochargersor superchargers, are called turbocharged engines. Figure 19 is an illustration of aturbocharged engine [Obert 1968].

Figure 19. Illustration of a Turbocharged Engine

Charge Air Cooling: Turbocharging or supercharging engines leads to increasingcharge air temperature and reducing its density. To improve charging of the combustion

chamber, designers resort to cooling the charging after it had been boosted (increasingits pressure), hence the term aftercooling is used to describe a system through which airpasses through a charge cooler, as illustrated in Figure 19. In the aftercooler heat isexchanged to another medium, normally water or air, prior to its introduction into the




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cylinder. Engines are often described as turbocharged and aftercooled (TA). Somedesigners refer to this heat exchanger as an intercooler since it is placed between theturbocharger and the intake manifold of an engine or port of a cylinder.

Engine Cooling: Combustion produces heat that must be dissipated away from theengine hottest regions to preserve its mechanical integrity. The process of carrying heataway from the engine's critical component is achieved through its cooling system.

While most engines use a water jacket around the hot engine parts to transfer the heatfrom the engine to the environment via a heat exchanger (radiator), some engines useair cooling. The liners in these engines have fins on their outer surface that exchangeheat by means of air blowing across those fins.

Power Modulation: Increasing or decreasing the engine's power output generallyrequires controlling its air-to-fuel ratio. In gasoline, spark-ignited engines powermodulation is achieved through throttling the engine, thus reducing its air flow andenriching its fuel and air mixture. Some engineers refer to this method as qualitativecontrol since it involves changing the quality of the fuel/air mixture from lean to rich orvice versa. On the other hand, compression ignition engines have an overall leanmixture at all operating conditions. They are referred to as quantitatively controlled

engines since power is controlled by varying the amount of fuel in an engine where airflow is mostly unchanged for a fixed speed.

There are several other designations that could be considered categories of engines, but they are a bitmore technically involved and used by those who are well versed with engine designs. Thesedesignations may include the type of combustion chamber used in an engine, such as M.A.N chamberfor a system having a design specific to a German company (shown in "The Case..."), or Quadram fora system designed by Perkins. Other designation may refer to the method of mixture preparation, andso on. Yet, the preceding classification or categorization covers most of the designs available.

Assessment of Engine PerformanceRegardless of categories, design features, importance of application, displacement, or any otherdescriptive feature engines must perform a certain task. It is reasonable to expect engines to givemaximum performance at minimum cost. In other words, what return should customers expect ontheir capital investment (price paid for the engine and its installation), and their operating cost (costof running and maintaining the engine). In addition and in view of current and future environmentalconcerns, it is not enough to maximize performance and minimize cost, but to do so while preservingthe environment.

Figure 20. Schematic Illustration of Control Volume (Mass Conservation)

It is perhaps time that we investigate engine performance through applying thermodynamic

principles. This material is presented in a simplified manner to facilitate it understanding especially tothose not familiar with this science. The fundamental reason for resorting to thermodynamics is that itis an elegant way to treat the balance of energy as well as mass of the working fluid in a controlledvolume. The controlled volume in this case is the cylinder where energy and mass flow into and out




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of. Figure 20 is an illustration of the control volume in an engine where both chemical energy in theform of fuel as well as mass in the form of fuel and air are introduced.

Fuel is usually metered into the control volume where air is also inducted. Both fuel and air areintroduced into the cylinder in amounts (mass) that support efficient combustion and produce thedesired power output. In summary, the control volume receives mass and energy and also deliversmass and energy. Applying the laws of thermodynamics helps in quantifying the relationship between

the energy delivered to that received by the control volume.

The First and Second Laws of Thermodynamics

The first law of thermodynamics simply states that energy cannot be created nor destroyed but canonly be converted from one form to another. For instance, a mass of hydrocarbon fuel containschemical energy that is converted to work or mechanical energy within the cylinder (control volume)of an internal combustion engine. Theoretically, if this process was ideal and no losses are incurredthis energy conversion would be 100% efficient. Yet, in reality converting energy from one form toanother involves many losses resulting in an overall loss in efficiency. This fact is what the secondlaw of thermodynamics expresses as it states that the useful work from a combustion system should

be less than the energy input [Henein 1985]. In general, the ratio between useful work and the thermalenergy added to the control volume represents the brake thermal efficiency of the system.Considering the cylinder and piston arrangement shown in Figure 20, the combustible mixture of fueland air is burned in the control volume producing heat that results in the expansion of that volumecausing the piston to move. Motion of the piston creates friction against the cylinder walls leading tofriction heat loss. Another source of loss results from the temperature associated with the heatgenerated by the combustion process itself. As the combustion temperature increases, the cylindermaterial approaches its limitation in mechanical strength. Therefore, cylinders or control volumes arecooled, by water or air, to move heat away from the material thus preserving its mechanical strength.Heat transferred away from the control volume material is another loss added to the balance betweenthe energy received by that control volume and the energy it delivers back. Another major source ofloss in this energy conversion system is exhaust gases flowing out of the control volume. Exhaustgases exit the control volume with heat energy delivered to ambient without any benefit. They also

exit with a great deal of potential energy (pressure) as well as kinetic energy (speed). Therefore,considering the system on hand we can think of a control volume where fuel and air are supplied andin return piston motion (work) is delivered, but the work delivered is much less than the value ofenergy supplied to the control volume. The difference between the energy supplied to the controlvolume and that it delivers is the sum of losses including, but not limited to, cooling and exhaustlosses. A helpful illustration of this balance of energies is given in Figure 21.

Figure 21. Energy Balance in a Control Volume

While the description given so far for the energy exchange through a control volume is generallycorrect it may not be very complete. A better accounting of all the energies entering and leaving thecontrol volume will have to include other types of energies yet unaccounted for. For instance, any

mass entering the control volume brings with it several forms of energy [Van Gerpen 2000]:

internal energy; mainly due to its temperature which is generally very small




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kinetic energy; mainly due to injection characteristics which usually leads to importantinteractions between the fuel and air within the control volume

potential energy; generally associated with pressure admitting mass into the control volume

flow energy; principally associated with the inter-relation between the control volume and itspressure.

Revisiting the first law of thermodynamic and considering the various forms of energy we are nowacquainted with, one could think of the energy balance in a control volume as follows:

Net Output = (Energy Supplied to the Control Volume) - (Total Energy Loss)

In other words, a careful accounting of the energies supplied to the control volume and thosedelivered by the same volume can assist in assessing the system conversion efficiency. Caution must

be exercised when defining the control volume. So far, we have considered one cylinder in an engineas the control volume. However, an entire engine or the total vehicle can be viewed as that controlvolume. Therefore, defining the control volume and its boundaries is extremely important in theenergy conservation equation.

Mass Conservation in Combustion Thermodynamics

Combustion Mass Balance

To complete the discussion regarding the control volume, consideration should be given to the massentering and leaving it. As illustrated in Figure 20, fuel and air are the two principal constituentsentering the control volume. In view of modern diesel engine technology this may be a limitedtreatment of the masses entering the control volume since the need to control emission may dictaterecirculating some exhaust products back into the cylinder. In addition, lube oil contribution toexhaust emissions is being scrutinized and efforts are made to limit its consumption within thecylinder. Therefore, a more accurate representation of the mass balance in a control volume may bedescribed by the following relationship:

Exhaust Mass = Mass of Air + Mass of Recirculated Exhaust + Mass of Fuel +Mass of Lube Oil

or

Mexh = Ma + Mf+ Megr+ Mlo

In practical embodiments of reciprocating internal combustion engines accessories are used toperform various functions in support of the engine operation. For example, lubricating the enginerequires a oil pump driven by the engine itself, thus subtracting a portion of the work produced bythat engine. Therefore, if the control volume is the cylinder then the conversion efficiency of such

system will be higher that if the control volume was an engine. In fact, in diesel engines power isconsumed in driving its fuel injection system. Other drives that are required for the engine operationinclude the camshaft, alternator, coolant pump, and superchargers.

Diesel Fuel Composition

Even though combustion in diesel engines is not the subject of this discussion, an alternativetreatment of the mass balance in a control volume, cylinder in this case, can be helpful. A simplifiedexpression for mass conservation can be limited to reactant species and reaction products such as fuel(CnHm) and air (mostly nitrogen and oxygen) reacting with each other in the proper environment and

producing exhaust constituents as follows:

CnHm + aO2 + 3.76aN2 bCO2 + cCO + dH2O + eOH2 + fH2 + g(HC) +

hNO + iHCHO + jNH3 + kN2

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In Equation 13, CnH

mrepresents hydrocarbon fuel reacting with air consisting of oxygen and

nitrogen in a volumetric ratio of 20.99 to 79.01%, respectively. It follows that for every mole ofoxygen provided, 3.76 moles of nitrogen would be present in the reaction. The majority of theexhaust gaseous species are CO

2, H

2O, N

2, and in diesels excess O

2as well. In fact, these species

constitute about 99% of the exhaust from an engine, leaving only 1 percent unaccounted for but madeof mostly undesirable species.

Based on a measured average molecular mass of diesel fuel and its carbon to hydrogen ratio, one cancalculate the average chemical formula for the diesel fuel. The following calculation is based on themolecular mass of diesel fuel of 191 (as determined by UOP Method 375-86 [Van Gerpen 2000]). Sincethe molecular mass of C is 12.0111 and that of hydrogen is 1.00797, the hydrocarbon designation ofdiesel fuel can be determined as follows:

12.0111n + 1.00797m = 191

From actual fuel analysis each kg of diesel fuel contains 0.8616 kg of carbon or:

0.8616 kg C / 12.0111 = 0.07173 kmol C

Similarly

0.1251 kg H / 1.00797 = 0.12411 kmol H

From Equations 15 and 16, the hydrogen to carbon ratio

m/n = H/C = 0.12411 / 0.07173

By solving Equations 14 and 17, we can define diesel fuel as C13.883

H24.053

. Since diesel fuels are

mixtures of hydrocarbons of variable composition, a certain variability of the above carbon and

hydrogen designations will be seen in real life samples.

Stoichiometric Ratio in Diesel Combustion

Having determined the composition of diesel fuel it is relatively simple to calculate its stoichiometricratio. By definition, the stoichiometric ratio is the ratio of air to fuel that when fully combusted wouldyield nothing but CO

2, H

2O, and N

2. It is sometimes referred to as the chemically correct ratio.

Applying this definition to the diesel fuel from the previous section yields the following:

C13.883H24.053 + 94.744[0.21 O2 + 0.79 N2] + 13.883 CO2 + 12.026 H2O +

74.848 N2

From Equation 18, the molar A/F ratio is 94.744 kmol air/kmol fuel. On a mass basis the A/F ratiocan be calculated as follows:

[94.744 kmol air/kmol fuel] [28.97 kg air/kmol air] [kmol fuel/191 kg fuel]= 14.37 kg air/kg fuel

Additional Performance Parameters and TheirDefinitionAfter months and years designing an engine it must be put to the test of verifying whether or not alldesign assumptions and choices can indeed perform the intended output in an efficient manner. To

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fully appreciate what an engine can do the engineer must be well versed with the science of testingengines including laboratory tools designed for this purpose, evaluation and performance parameters,test methods, interpreting results, statistical tools, equipment calibration, and maintenance of test andanalysis equipment. Of importance to the end user is the following:

The engine acquisition cost

The engine operating characteristics (output and speed)

The engine operating cost (fuel consumption and maintenance)

The engine reliability and its durability

The engine exhaust emission profile and the cost of making it environmentally friendly.

Of course, the importance of any of these items will depend on the specific application of the engine.Therefore, priorities have to be established before acquiring an engine to reflect the purpose forwhich the engine will be used. Nevertheless, there are common methods that are used to evaluateengine performance as well as their emission profiles. In this section, engine performance parameters,test tools, and test methodologies will be described and in some cases quantified.

Pow er Output

Of primary importance to the designer, engine development engineer, end user, and others in theindustry is the power output of an engine. The power output is defined by a maximum torque at agiven engine speed. The engine maximum power output, sometimes referred to as rated power, isdefined as follows:

Power Output = (T N) / Constant

where T is engine torque in lb-ft or N.m, and N is engine speed in rotations per minute. Theappearance of torque in Equation 20 leads us to seek a more basic definition of power by consideringthe fundamental purpose for using an engine. An engine produces work that we can use in various

applications, and power is the rate at which this work is produced. The most common method used tomeasure power is a device designed to brake the engine and is called a dynamometer. There are manytypes of dynamometers, but they have certain features in common. Most dynamometers are designedwith a stator that does not rotate and is coupled electromagnetically to a rotor as shown in Figure 22.The rotor is the second major feature of a dynamometer and is concentric with the stator. It is alsocoupled to the engine and rotates around the same axis as the stator. In electromagneticdynamometers, the rotor is driven by the engine and an electric field in the stator tries to oppose itsmotion. The electromagnetic force (F) exerted is measured by a load cell that is placed at distance (b)from the center of the load cell as shown in Figure 22. The product of the force (F) and distance (b) asexpressed in Equation 21 defines torque.

Figure 22. Schematic Diagram of an Engine Dynamometer

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T = F b

In one revolution the rotor travel a distance 2 r against resistance force f, thus its work for this onerevolution is:

Work = 2rf

The rotor turning moment (rf) must be balanced exactly by the external turning moment that is theproduct (Fb) and this relationship can be expressed as follows:

Rf = Fb

Substituting from Equation 23 into 22 we can then express work as:

Work = 2Fb

Using the relationship of Equation 21 to express work per unit time gives the following:

Work per minute = 2TN

Since power is defined as the rate of doing work, we can then write the following:

Work per minute = Power = 2TN

Horsepower (hp) is a unit of power equal to 33,000 lb-ft per minute (550 lb-ft per second) in theEnglish system, and the kilowatt (kW) is its equivalent in the metric system. The kilowatt is 550 0.746 = 738 lb-ft per second. Having defined these units, we can then express horsepower as:

hp = 2TN / 33,000 = T N / 5252

Equation 27 establishes a relationship between power output, torque, and engine speed. From thisrelationship we conclude that since hp is a function of both torque and speed, then it is possible todesign an engine to achieve power through high torque or through high speed. Engines designed forhigh torque output are usually large and built to withstand high internal forces. Their maximumspeeds are quite low ranging from a few hundred revolutions per minute (rpm) to perhaps 1800 to2100 rpm. Stationary engines are on the low end of the speed spectrum while on highway trucks andmid-range engines occupy the high end of the speed range. High torque/low speed engines are verysuitable for heavy-truck applications enabling them to move extremely high loads away from loadingdocks or traffic lights at very low engine speeds. Meanwhile small high speed engines, such as those

powering many passenger cars, produce their high power at very high engine speeds (above 5,000rpm). An extreme example of these types of applications is the race car engine with speeds far

exceeding 9,000 rpm.

Indicated Pow er

The brake power, as discussed above, has been so named after the method used for its quantification.Brake power is a good measure for the useful power produced by the engine. Indicated power is thatwhich is produced by the direct application of the gas pressure on the surface of the piston. Byintegrating cylinder pressure during a complete combustion cycle one can obtain an value for thecycle pressure that can be used to calculate indicated power.

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Figure 23. Pressure-Volume Diagram

From Figure 23 the net pressure applied to the surface of the piston is (Area A - Area B). Area A is

bordered by the clockwise arrows and is regarded as positive pressure on the piston. This is aconventional way rather than a purely scientific way since during the compression stroke, it is thepiston that is doing work on the gas. A similar consideration is given to Area B, where the piston isfound once again to be pushing exhaust gases while it moves from BDC to TDC. Area B is oftenconsidered a measure for pumping losses. In addition to the Area B, friction losses are also subtractedfrom indicated pressure to obtain brake power. In other words:

Indicated Power = Brake Power + Pumping Power + Friction Power

In turbocharged engines pumping losses are relatively small and can be neglected and resulting in thefollowing relationship:

Indicated Power = Brake Power + Friction Power

Mechanical Efficiency

The term friction power in Equation 29 includes power required to expel exhaust gases, induct freshair, overcome piston ring/liner friction, bearing friction, and drive engine accessories as well asaccount for parasitic losses elsewhere within the engine. It follows that mechanical efficiency (mech)

can be defined as the ratio of useful power (brake power) over available power (indicated power). Inmathematical form it can be expressed as follows:

Mechanical Efficiency = (Brake Power) / (Indicated Power)

or

mech = Pb/Pi = 1 - (Pf/Pi)

where Piis indicated power, P

bis brake power, and P

fis friction power. Normally friction power

losses increase with engine speed. An estimate of friction power can be obtained from normal enginebrake power and fuel consumption basic relationship. An example is given in Figure 24 where regularengine dynamometer test information of brake horsepower and fuel consumption can be used atconstant engine speed to derive an estimate of friction power. By plotting brake power on the X-axisand fuel consumption on the Y-axis, data from zero to 300 horsepower are represented by an almost

straight line except for the high output points. As this line is extended to the left and intersects the X-axis it gives us an estimate of the friction power as shown in Figure 24. At the point where that lineintersects the vertical line at zero power output, the fuel consumption required to overcome enginefriction at that speed can be read (about 15 lb/hr in this case). The line describing engine brake power

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versus fuel consumption, its extension to the left, shown in dotted line, and the estimate of frictionfuel consumption were attributed to an Englishman by the name of Willans, and hence the namegiven to this characteristic line.

Even though diesels are not usually throttled, operating them at high speed and requiring greatervolume of air to be inducted into the cylinder through the narrow passage of an intake valve causes

pumping losses to increase and volumetric efficiency to degrade. For this as well as other reasons, in

modern engines valve area was increased by adopting two intake valves per cylinder instead of justone.

Figure 24. The Willans Line

Indicated Mean Effective Pressure

Area A in Figure 23 was described as the area indicating the work exerted by the working fluid on thesurface of the piston. Pressure resulting from the combustion process, as a function of the cylindervolume, is applied to the piston surface to produce power. The indicated mean effective pressure(imep) is the work exerted by the gas on the piston per unit swept volume. The following is anexpression of imep:

imep =(Pi/Vd) dV

Graphically, indicated mean effective pressure is illustrated in Figure 25.

Figure 25. Graphical Representation of Indicated Mean Effective Pressure

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Brake Mean Effective Pressure

While imep is a form of the bulk pressure exerted on the surface of the piston, brake mean effectivepressure may be described as the portion of that imep that produces useful power. It represents the network after subtracting friction, pumping, and other parasitic losses. The largest of these losses beingthat of friction as expressed by in Equation 29. There is a relation of proportionality between brake

mean effective pressure (bmep) and engine torque. Rather than using torque which depends on enginesize, bmep is preferred since it is a normalized parameter as given in Equation 33.

bmep =(Pb/Vd) dV

Other useful relationships for bmep include those given in Equations 31 and 32. Numerically, bmepcan be calculated using its proportional relationship to torque as shown in Equations 34 and 35 fortwo- and four-stroke engines, respectively [Obert 1968].

bmep = 75.4 T / CID, psi

bmep = 150.8 T / CID, psi

where CID is the engine displacement in cubic inch. It is customary to compare engines on the basisof their bmep using their peak torque value. However, a more accurate comparison of engines shouldconsider the entire bmep characteristic of engines versus their speed range.

Specific Fuel Consumption

Fuel consumption in mass per unit time is normally recorded during engine testing. One way toevaluate engine efficiency is to ratio fuel consumption to useful power. The result of thismathematical treatment produces the term brake specific fuel consumption (bsfc). In essence, bsfc isa measure of how much fuel an engine consumes in the process of producing an output of one

horsepower (see Equation 36).

bsfc = (Fuel Consumption per Unit Time) / (Brake Power Output), lb/bhp-hr

Single cylinder engines are popular in engine research work. Since power is produced by one cylinderonly, friction and parasitic losses are disproportionately high relative to the power output of theseengines. In these cases, rather than using the term bsfc engineers compute indicated specific fuelconsumption (isfc). The term isfc is calculated on the basis of indicated power output as given inEquation 37.

isfc = (Fuel Consumption per Unit Time) / (Indicated Power Output), lb/bhp-hr

Volumetric Efficiency

Volumetric efficiency is a measure of the breathing quality of an engine. It is the ratio of the mass ofair actually inducted by the engine during its intake stroke to the theoretical mass of air that could beinducted given the displacement of that engine. Equation 38 puts this relationship into a mathematicalform.

vol= (Actual Mass of Air Inducted) / (Theoretical Mass of Air to Fill

Displacement)

Obert [Obert 1968] describes this definition as a misnomer because it is a ratio of two masses rather two

volumes. Yet, it is the conventional way of calculating volumetric efficiency since it takes air densityinto consideration.

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Engine Specific Weight and Volume

Total engine weight and volume are very important for packaging and cost considerations. Forinstance, under-hood space for many on highway as well as off road applications is very limited.Therefore, engines having the right power output and torque characteristics may have differentvolumes where the smallest volume would be of greatest advantage. Not only is engine volume very

important, but also its weight. The lighter the weight the greater its advantage for a given applicationsince it cost fuel to transport additional weight. The terms specific engine weight and specific enginevolume are used to compare engines of similar power outputs, but having different size and weight.Equations 39 and 40 give the calculation for specific volume and specific weight, respectively.

Specific Volume = (Engine Volume) / (Rated Power)

Specific Weight = (Engine Weight) / (Rated Power)

ReferencesBlair, G.P., 1996. "Design and Simulation of Two-Stroke Engines", The Society of Automotive Engineers, Warrendale, PA

Foulkes, D.M., 1995. "Developing Light-Duty Diesel Engines For Low Emissions and High Fuel Economy", Internal

Report, Ford Motor Company

Henein, N., 1985. "Engine Fundamentals", Lecture Notes to Ford Tractor Operations, Ford Motor Company, October 16 -

November 20, 1985

Heywood, J.B., 1988. "Internal Combustion Engine Fundamentals", McGraw-Hill, New York

Obert, E.F., 1968. "Internal Combustion Engines", International Text Book

Power, 1964. "Oil and Gas Engines", Special Report, Power, 330 West 42nd St. New York

Taylor, C.F., 1985a. "The Internal Combustion Engine in Theory and Practice", M.I.T. Press, Volume 1, Revised Edition

Van Gerpen, J., 2000. "The Origins of Fuel Economy", Diesel Engine Technology Engineering Academy, SAE EngineeringAcademies

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