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In the diesel engine, ignition of the fuel is accomplished by the heat of compression alone. To support combustion, air is required. Approximately 14 pounds of air are required for the combustion of 1 pound of fuel oil. However, to insure complete combustion of the fuel, an excess amount of air is always supplied to the cylinders. The ratio of the amount of air supplied to the quantity of fuel injected during each power stroke is called the air-fuel ratio and is an important factor in the operation of any internal-combustion engine. When the engine is operating at light loads there is a, large excess of air present, and even when the engine is overloaded, there is an excess of air over the minimum required for complete combustion.

The injected fuel must be divided into small particles, usually by mechanical atomization, as it is sprayed or injected into the combustion chamber. It is imperative that each of the small particles be completely surrounded by sufficient air to effect complete combustion of the fuel. To accomplish this, the air in the cylinder must be in motion with good fuel atomization, combined with penetration and distribution. In mechanical injection engines this is accomplished by forcing scavenging air into the cylinder with a whirling motion to create the necessary turbulence. This is usually done, in the 2-cycle engine, by shaping the intake air ports, or by casting them so that their centers are slightly tangential to the axis of the cylinder bore.

Before proceeding with the study of the combustion process, the conditions considered essential to good combustion should be reviewed:

2. The fuel must enter the cylinder in a fine mist or fog.

3. The fuel must mix thoroughly with the air that supports its combustion.

4. Sufficient air must be present to assure complete combustion.

5. The temperature of compression must be sufficient to ignite the fuel.

Figure 9-1 is a reproduction of a pressure-time diagram of a mechanical injection engine. The lower curvy part of which is a dotted line, is the curve of compression and expansion when no fuel is injected. At A the injection valve opens, fuel enters the combustion chamber and ignition occurs at B. The pressure from A to B should fall slightly below the compression curve without fuel due to absorption of heat by the fuel from the air. The period from A to B is the ignition delay. From B the pressure rises rapidly until it reaches a maximum at C. This maximum, in some instances, may occur at top dead center. At D the injection valve closes, the fuel is cut off, but burning of the fuel continues to some undetermined point along the expansion stroke.

The height of the diagram from B to C is called the firing pressure rise and the slope of the curve between these two points is the rate at which the fuel is burned.

Poor combustion of the fuel is usually indicated by a smoky exhaust, but some smoke may be the result of burning lubricating oil that has passed the rings into the combustion chamber. Incomplete combustion is indicated by black smoke, or if the fuel is not igniting, it may appear as blue smoke. Immediately after starting an engine, when running at light loads or at overloads, or when changing from one load to another, smoke is likely to appear.

A smoky exhaust from the engine does not indicate whether one or all the cylinders are

9A2. Engine losses. It is obvious that not all of the heat content of a fuel can be transferred into useful work during the combustion process. The many different losses that take place in the transformation of heat energy into work may be divided into two classes, thermodynamic and mechanical. The net useful work delivered by an engine is the result obtained by deducting the total losses from the heat energy input.

Thermodynamic losses are caused by:

1. Loss to the cooling system and losses by

2. Heat rejected and lost to the atmosphere in the exhaust.

3. Inefficient combustion or lack of perfect combustion.

A loss due to imperfect or incomplete combustion is an important item, because such losses have a serious effect on the power that can be developed in the cylinder as shown by the pressure-volume diagram or indicator card. Complete combustion is not possible in the short time permitted in modern engine design. However, these losses may be kept to a minimum if the engine is kept adjusted to the proper operating condition. Incomplete combustion can frequently be detected by watching exhaust temperatures, noting the exhaust color, and being alert for unusual noises in the engine.

Heat energy losses from both the cooling water systems and lubricating oil system are always present. Some heat is conducted through the engine parts and radiated to the atmosphere or picked up by the surrounding air by convection. The effect of these losses varies according to the part of the cycle in which they occur. The

1. Heat losses to jackets during compression, combustion, and expansion phases of the working cycle.

2. Heat losses during the exhaust stroke.

3. Heat losses absorbed by the walls of the exhaust passages.

4. Heat generated by piston friction on cylinder walls.

Heat losses to the atmosphere through the exhaust are inevitable because the engine cylinder must be cleared of the still hot exhaust gases before another fresh air charge can be introduced and another power stroke begun. The heat lost to the exhaust is determined by the temperature within the cylinder when exhaust begins. It depends upon the amount of fuel injected and the weight of air compressed within the cylinder. Improper timing of the exhaust valves, whether early or late, will result in increased heat losses. If early, the valve releases the pressure in the cylinder before all the available work is obtained; if late, the necessary amount of air for complete combustion of the next charge cannot be realized, although a small amount of additional work may be obtained. The timing of the exhaust valve is a compromise, the best possible position of opening and closing being determined by the engine designer. It is essential that the valve be tight and properly timed in order to maintain the loss to the exhaust at a minimum. This is also true for air inlet valve setting on 4-cycle type engines.

If an indicator card is taken of a diesel engine cylinder, it is possible to calculate the horsepower developed within the cylinder. This calculation does not take into account the power loss resulting from mechanical or friction losses, as will be discussed later, but it reflects the actual work produced within the cylinder.

Mechanical losses are of several kinds, not all of them present in every engine. The sum total of these mechanical losses deducted from the indicated horsepower developed in the cylinders will give the brake horsepower finally delivered as useful work by the engine. These mechanical or friction losses include bearing friction, piston and piston ring friction, and

pumping losses caused by operation of water pumps, lubricating oil pumps, and scavenging air blowers, power required to operate valves, and so forth. Friction losses cannot be eliminated, but they can be kept at a minimum by maintaining the engine in its best mechanical condition. Bearings, pistons, and piston rings should be properly installed and fitted, shafts must be in alignment, and lubricating and cooling systems should be at their highest operating efficiency.

9A3. Compression ratio and efficiencies. a. Compression ratio. The term compression ratio is used quite extensively in connection with engine performance and various types of efficiencies. It may be defined as the ratio of the total volume of a cylinder to the clearance volume of the cylinder. It may be best explained by reference to the pressure-volume indicator card of a diesel cylinder. In Figure 9-3, the volume is reduced from square root(C) + square root(D) to square root(C) during compression. The compression ratio is then equal to
(square root(C) + square root(D))/square root(C)

Compression ratio influences the thermal efficiency of an engine. Theoretically the thermal efficiency increases as the compression ratio is increased. The minimum value of a diesel engine compression ratio is determined by the compression required for starting, which, to large extent is dependent on the type of fuel used. The maximum value of the compression ratio is not limited by the fuel used but is limited by the strength of the parts of the engine and the allowable engine wgt/bhp output.

b. Cycle efficiency. The efficiency of any cycle is equal to the output divided by the input. The diesel cycle shows one of the highest efficiencies of any engine yet built because of the higher compression ratio carried and because of the fact that combustion starts at a higher temperature. In other words, the heat input is at a higher average temperature. Theoretically, the gasoline engine using the Otto or constant volume cycle would be more efficient than the diesel if it could use compression ratios as high as the latter.

The gasoline engine operating on the Otto cycle cannot use a compression ratio comparable to the diesel engine due to the fact that the fuel and air are drawn in together and compressed. If high compression ratios were used,

The temperature-entropy (T-S) diagram of any particular cycle indicates the amount of heat input and the amount of heat rejected. For example, in Figure 9-4, the T-S diagram of a modified diesel cycle, the heat input is represented by the area FBDG and the heat rejected to the exhaust by the area FAEG. The heat represented in doing useful work is represented by the difference between these two, or area ABDE. The efficiency of the cycle can then be expressed as
(H₁-H₂)/H₁ where H₁ is the heat input along lines BC and CD (the lines representing the constant volume and constant pressure combustion), and H₂ is the heat rejected along line EA (the line representing the constant volume exhaust). Since heat and temperature are proportional to each other, the cycle efficiency is actually computed from measurements made of the temperature. The specific heat of the mixture in the cylinder is either known or assumed, and when combined with the temperature, the heat content can be calculated at any instant. Thus, it is seen that temperature is a measure of heat, and that the heat is proportional to the temperature of the gas.

c. Volumetric efficiency. The volumetric efficiency of an engine is the ratio of the volume that would be occupied by the air charge at atmospheric temperature and pressure to the cylinder displacement (the product of the

Figure 9-4. Temperature-entropy diagram of modified diesel cycle.

Volumetric efficiency is actually the completeness of filling of the cylinder with fresh air at atmospheric pressure. The volumetric efficiency of an engine may be increased by enlarging the areas of intake and exhaust valves or ports, and by having all valves properly timed so that as much air as possible will enter the cylinders. Since any burned gases will reduce the charge of fresh air, the supercharging effect gained by early closing of the exhaust valves or ports will reduce the volumetric efficiency. In some engines, the volumetric efficiency is also increased by using special apparatus to utilize air at 2 to 3 psi over the atmospheric pressure. This procedure is commonly called supercharging.

d. Thermal efficiency. Thermal efficiency may be regarded as a measure of the efficiency and completeness of combustion of the injected fuel. Thermal efficiencies are generally considered as being of two kinds, indicated thermal efficiency and over-all thermal efficiency.

If all the potential heat in the fuel were delivered as work, the thermal efficiency would be 100 percent. This is not possible in practice, of course. To determine the values of the above efficiencies the amount of fuel injected is known, and from its heating value, or Btu per pound, the total heat content of the injected fuel can be found. From the mechanical equivalent of heat (778 foot-pounds are equal to 1 Btu), the number of foot-pounds of work contained in the fuel can be computed. If the amount of fuel injected is measured over a period of time, the rate at which the heat is put into the engine can be converted into potential power. Then, if the indicated horsepower developed by the engine is

Indicated thermal efficiency =
(Indicated hp X 42.42 Btu per minute per hp) / (Rate of heat input of fuel in Btu per minute) X 100 percent

In like manner the over-all thermal efficiency can be found from the brake horsepower or the actual power available at the engine shaft.*

Over-all thermal efficiency =
Brake horsepower / Heat input of fuel X 100 percent

e. Mechanical efficiency. The mechanical losses in an engine decrease the efficiency of the engine and represent the skill with which the engine parts were designed as well as the skill with which the operator maintains the engine. As previously stated, the brake horsepower is equal to the indicated horsepower minus the mechanical losses. The ratio of brake horsepower to indicated horsepower, then, is the mechanical efficiency of the engine which increases as the mechanical losses decrease.

Mechanical efficiency =
Brake horsepower / Indicated horsepower X 100 percent

b. Fuel characteristics. The cetane number of the fuel has an important effect on engine performance. Fuels with low cetane rating have high ignition lag. A considerable amount of fuel collects in the combustion space before ignition occurs, with the result that high maximum pressures are reached, and there is a tendency toward knocking. This tends to increase wear of the engine and reduce its efficiency. Fuels with high cetane ratings have low auto-ignition temperatures and hence are easier starting than fuels with low cetane ratings. Therefore, diesel engine performance is improved by the use of high cetane number fuel oils.

c. Air temperature. The temperature of the air in the cylinder directly affects the final compression temperature. A high intake temperature results in decreased ignition lag and facilitates easy starting, but is generally undesirable because it decreases the volumetric efficiency of the engine.

d. Quantity of fuel injected per stroke. The quantity of fuel injected determines the amount of energy available to the engine, and also (for a given volumetric efficiency) the air-fuel ratio.

e. Injection timing. The injection timing has a pronounced effect on engine performance. For many engines, the optimum is between 5 degrees to 10 degrees before top dead center, but it varies with engine design. Early injection tends toward the development of high cylinder pressures, because the fuel is injected during a part of the cycle when the piston is moving slowly and combustion is therefore at nearly constant volume. Extreme injection advance will cause knocking. Late injection tends "to decrease the mean indicated pressure (mip) of the engine and to lower the power output. Extremely late injection tends toward incomplete combustion, as a result of

f. Injection rate. The rate of injection is important because it determines the rate of combustion and influences engine efficiency. Injection should start slowly so that a limited amount of fuel will accumulate in the cylinder during the initial ignition lag before combustion begins. It should proceed at such a rate that the maximum rise in cylinder pressure is moderate, but it must introduce the fuel as rapidly as permissible in order to obtain complete combustion and maximum expansion of the combustion products.

g. Atomization of fuel. The average size of the fuel particles affects the ignition lag and influences the completeness of combustion. Small-sized particles are desirable because-they burn more rapidly. Opposed to this requirement is the fact that small particles have a low penetration, and there is therefore a tendency toward incomplete mixing of the fuel and the combustion air, which leads to incomplete combustion.

h. Combustion chamber design. The amount of turbulence present in the combustion chamber of an engine affects the mixing of the fuel and the combustion air. High turbulence is an aid to complete combustion.

9B2. Power. Engine performance of an internal-combustion engine may be measured in terms of torque, or power developed by the engine. The power that any internal-combustion engine is capable of developing is limited by mean effective pressure, length of stroke, cylinder bore, and the speed of the engine in revolutions per minute (rpm).

a. Mean indicated pressure. The average or mean pressure exerted on the piston during each expansion or power stroke is known as the mean indicated pressure. Mean indicated pressure is of great importance in engine design. It can be obtained from indicator cards mathematically or directly from the planimeter. Excessive mean pressures result in overloading the engine and consequent high temperatures. Temperatures greater than those contemplated in the engine design may cause cracked cylinder heads, liners, and warped valves. There are two kinds of mean effective pressures. One, mip, or mean

NOTE. Maximum pressure developed has no bearing on mep.

b. Length of stroke. The distance the piston travels from one dead center to its opposite dead center is known as the length of stroke. This distance is one of the factors that determines the piston speed which is limited by the frictional heat generated and the inertia of the moving parts. In modern engines, piston speed reaches approximately 1600 feet per minute. If the length of stroke is too short, excessive side thrust will be exerted on a trunk type piston. The length of stroke, however, cannot be too great because of the lack of overhead space available on submarine type engines.

c. Cylinder bore. The cylinder bore is its diameter, and from this the cross-sectional area of the piston is determined. It is upon this area that the gas pressure acts to create the driving force. This pressure is the mean indicated pressure referred to above, expressed and calculated for an area of 1 square inch. The ratio of length of stroke to cylinder bore is somewhat fixed in engine design. There are a few instances in which the stroke has been less than the bore, but in almost every case the stroke is longer than the bore. This ratio in a modern trunk-piston type engine is about 1.25, while in a crosshead type engine in use today it is about 1.50.

d. Revolutions per minute. This is the speed at which the crankshaft rotates, and since the piston is connected to the shaft, it determines, with the length of stroke, the piston speed. Since the piston moves up and down each revolution, the piston speed is equal to twice the stroke times the revolutions per minute (rpm), and is usually expressed in feet per minute. If the stroke is 10 inches, and the speed of rotation is 750 rpm, the piston speed is

750 X 2 X (12/10) = 1,250 feet per minute.

The power developed by the engine depends upon the engine's speed and the type of engine. If it is a single-acting, 4-stroke cycle engine there will be one power stroke for every two revolutions of the crankshaft. If it is a

Having defined the factors influencing the power capable of being developed, the general formula for calculating horsepower is as follows:

IHP = (P X L X A X N) / 33,000

P = Mean indicated pressure, in psi
L = Length of stroke, in feet
A = Effective area of the piston in square inches
N = Number of power strokes per minute

The horsepower developed within the cylinder as a result of combustion of the fuel can be calculated by measuring the mean indicated pressure and engine speed. Then with the bore and stroke known, the horsepower can be computed for the type of engine being used. This power is called indicated horsepower because it is obtained from the pressure measured from an engine indicator card. It does not take into account the power loss due to friction, as will be discussed later. Example:

Given a 12-cylinder, 2-cycle, single-acting engine having a bore of 8 inches and a stroke of 10 inches. Its rated speed is 720 rpm. When running at full load and speed, the mean indicated pressure is measured and is found to be 105 psi. What is the indicated horsepower developed by the engine?

Solution:

From the formula

IHP = (P X L X A X N) / 33,000

P = 105
L = 10 / 12
A = 3.1416 (8/2)²
N = 720

IHP = (105 X (10 /12) X 3.1416 (8/2)²) X 720

IHP = 96.96

Since this is just the horsepower developed in one cylinder, if the load is perfectly balanced among all cylinders, the total indicated horsepower of the engine is

IHP = 12 X 96.96 = 1163.5

BHP = IHP minus the mechanical losses.

From the example above, the IHP was found to be 1163.5. If the brake horsepower of this engine was 900 as determined in a test laboratory, then the mechanical losses would be

1163.5 - 900 = 263.5 horsepower

or

(263.5 / 1163.5) X 100 = 22.6 percent of the indicated horsepower developed in the cylinders

or 90 / 1163.5 = 77.4 percent mechanical efficiency.

Engine power is frequently limited by the maximum mean pressure allowed. To find the bmep of the above engine, first obtain the power developed in one cylinder. Thus,

900 / 12 = 75.0 bhp

From the general formula for horsepower,

HP = (P X L X A X N) / 33,000

75 = P X (10/12) X 3.1416 X (8/2)² 720 /33,000

P = (75 X 33,000) / (10/12 X 3.1416 X (8/2)² X 720)

P = 82.1 psi

Hence, for the above engine under the conditions stated the bmep is 82.1 while the mip is 105 psi.

The brake horsepower is the power available at the engine shaft for useful work. Brake horsepower cannot usually be measured after an engine is installed in service, unless the engine drives an electric generator. The brake horsepower is determined by actual tests in the shops of the manufacturer before delivery of the engine. Frictional losses are quite independent of the load on the engine. Hence, unless the brake horsepower has been measured at various loads and speeds, the mechanical losses cannot

9B3. Engine performance limitations. The power that can be developed by a given size cylinder whose piston stroke is fixed is limited only by the piston speed and the mean effective pressure. The piston speed is limited by the inertia forces set up by the moving parts and the problem of lubrication due to frictional heat.

The mean indicated pressure is limited by:

1. Heat losses and efficiency of combustion.

2. Volumetric efficiency or the amount of air charged into the cylinder and the degree of scavenging.

3. Complete mixing of the fuel and air which requires fine atomization, sufficient penetration, and a properly designed combustion chamber.

The limiting mean effective pressures, both brake and indicated, are prescribed by the manufacturer or the Bureau of Ships and should never be exceeded. In a direct-drive ship, the mean effective pressures developed are determined by the rpm of the shaft. In electric-drive ships, the horsepower and mep can be determined readily from the electrical readings, taking into account generator efficiency.

The diesel operator should remember that the term overloading means exceeding the limiting mean effective pressure.

9B4. Operation. All submarine type diesel engines are rated at a given horsepower and a given speed by the manufacturer. These factors should ordinarily never be exceeded in the operation of the engine. Using the rated speed and bhp, it is possible to determine a rated bmep which each individual cylinder should never exceed, otherwise that cylinder will become overloaded. The rated bmep holds only for rated speed. If the speed of the engine drops down below rated speed, then the cylinder bmep which should not be exceeded generally drops down to a lower value due to propeller characteristics. The bmep should never exceed the normal mep at lower engine speed. Usually it

Navy type engines are generally rated higher for emergency use than would normally be the case with commercial engines. The economical speed for most Navy type diesel engines is found to be about 90 percent of rated speed. For this speed the optimum load conditions have been found to be from 70 percent to 80 percent of the rated load or output. Thus, we speak of running the engines at an 80-90 combination which will give the engine parts a longer life and will keep the engine itself much cleaner and in better operating condition. The 80-90 means that we are running the engine with 80 percent of rated load at 90 percent of rated speed.

Some engine manufacturers design their engine fuel systems so that it is impassible to exceed the rated bmep to any great extent. This is done by limiting the maximum throttle or fuel control setting by means of a positive stop. This regulates the maximum amount of fuel that can enter the cylinder and therefore the maximum load of the cylinder.

An overloaded condition of an engine, or of one or more of its cylinders, may be indicated by:

1. Black smoke in the exhaust.

2. High exhaust temperature.

3. High lubricating oil and cooling water temperature.

4. Hot bearings and high temperatures of other engine parts (in general, a hot running engine).

5. Excessive vibration of the engine.

6. Unusual sound of the engine.

When black smoke is observed in the exhaust from the mufflers, it is not possible to determine immediately whether the entire engine or just one of the cylinders is overloaded. However, by opening the indicator cocks on the individual cylinders, the color of their exhausts can be determined.

High temperatures of the exhaust gases

Thermometers are provided in the lubricating oil and cooling water systems. Modern diesel engines have thermometers installed in the cooling systems of individual cylinders. An abnormal rise in any of these temperatures may indicate an overloaded condition and should be investigated as quickly as possible.

In general, excessive heat in any part of the engine may indicate overloading. An overheated bearing may be the result of overloading a cylinder. An abnormally hot crankcase could result from overloading the engine as a whole. Excessive temperatures of some engine parts can be checked by touch.

If all cylinders are not doing an equal

9C2. Causes of unbalance. In the preceding section some of the general causes of unequal load distribution were discussed. To prevent unbalance in an engine, the foremost consideration is that the engine must be in excellent mechanical condition. A leaky valve or fuel injector, leaky compression rings, or any other such mechanical difficulties will make it impossible for the operator to balance the load unless he secures the engine and dismantles at least a part of it. Therefore, the engine must be placed in proper mechanical condition before the load can be balanced.

Since the heat of compression is relied upon to ignite the fuel injected in the diesel engine, the amount of this compression must be maintained within fixed limits. In order to have the same type of combustion in each cylinder, the degree of compression in all cylinders should be approximately the same. For example, low compression pressure in one cylinder may prevent all the fuel from burning, or may even prevent ignition of the fuel in that cylinder. This would result in a reduced amount of work or no work being done by this cylinder. The common causes of low compression are:

1. Sticking compression rings.

2. Excessive ring or cylinder wear.

3. Leaky cylinder head gasket or cracked cylinder head.

4. Leaky valve in cylinder head.

5. Cracked cylinder liner.

6. Excessive clearance volume.

In correcting these, it is generally necessary to replace the defective part. However, in some cases such as a sticking ring or valve, it is necessary only to clean the part and replace it. These cold compression pressures with fuel cut

In order to have the load equally distributed, each cylinder must receive the same amount of fuel. It is here that the effect of an improperly adjusted fuel pump is evident. A cylinder receiving more fuel than necessary for a given load will develop more power than required.

Any adjustment of the fuel pump must be undertaken only by a person thoroughly familiar with the type of pump being used. He should first determine beyond all doubt that the engine is in proper mechanical condition. A great many factors may cause the cylinder to fire unevenly. Some of these causes are a clogged or improperly timed fuel injection valve, improperly timed air intake or exhaust valve, air or water in the fuel system, improper rocker arm valve clearance, dirt or other foreign matter in the fuel oil which may be plugging up the strainers and filters, and any other factor that contributes to poor combustion. If a cylinder is firing incorrectly, always check the above conditions before making any adjustments to the fuel pump.

Changing the amount of fuel being delivered by adjusting the pump should be done only when it is certain that the cause of the trouble is in the pump. This point cannot be emphasized too strongly. For instance, if the failure of a cylinder to fire correctly was due to a clogged fuel injection valve tip and the operator increased the fuel supply to the cylinder with the intention of increasing the power developed by that particular cylinder, the increase in fuel might wash the valve clean and cause the cylinder to become badly overloaded from the excess fuel supplied. The correct procedure would have been to replace the clogged injection valve with a spare and to clean the one that was removed. The decrease in power delivered by a cylinder may also be due to some foreign matter under a valve or piston ring, and once cleared, the cylinder would become overloaded if the fuel supply had, in the meantime, been increased.

The operator who always maintains his plant in good mechanical condition will be required to make few, if any, adjustments to the

After an overhaul in which piston rings of cylinder liners have been renewed, considerable adjustment of the engine may be necessary. Lubricating oil will leak by the rings into the combustion space until after the rings have properly seated. The compression will also increase as the seal between the rings and the liner becomes more effective. The lubricating oil will burn in the cylinder, giving an incorrect indication of fuel oil combustion, and if the pump has been properly set when the engine was started, the engine will be overloaded, or at least unbalanced. As the compression rises to normal pressure, the power developed will increase as also will the conditions of pressure and temperature under which the combustion takes place. Hence, when an overhaul has been completed, the engine must be carefully watched until the rings are seated, and the compression set to the level specified in the instructions for that type of engine. This adjustment will be facilitated by the use of frequent compression tests. If the engine is not fitted so that the compression can be readily varied, the engine should be run under light load until it is certain that the rings have seated.

9C3. Effect of unbalance. In general, the effect of unbalance is an overheated engine. Clearances are established by the engine designer to allow for sufficient expansion of moving parts when operating at the designed temperatures. Consequently, an engine operating at temperatures in excess of those for which it was designed may suffer many casualties. Excessive expansion of the moving parts will cause seizures and a burning up of the engine. If the temperatures rise above the flash point of the lubricating oil vapors in the crankcase, an explosion may result. The high temperatures may destroy the lubricating oil film between adjacent surfaces of the moving parts and result in further increased temperatures due to the increased friction. In fact, the effect is the same as for overheating from any cause.

Since the mean indicated pressure

If an engine is developing 600 bhp, and its mechanical efficiency is 80 percent, the indicated horsepower being developed is 750. If the engine has 10 working cylinders, each cylinder should be producing 75 indicated horsepower. When this is not the case the engine is unbalanced. The effect here would be to increase the mean indicated pressure of those cylinders doing less than their share of the work, and to decrease that of those cylinders producing more than 75 indicated horsepower.

The turning moment acting on the crankshaft is proportional to the force acting on the piston. This force, in turn, is the result of the mean indicated pressure developed in the cylinder. If these forces from different cylinders are not equal, there is an uneven turning moment acting along the length of the crankshaft, and vibrations result. These vibrations, if sufficiently severe, may shake the engine loose in its foundation, crack the engine housing, framework, and bedplate, destroy the bearings, and even break the crankshaft. It is obvious that a badly vibrating engine can result in serious damage and should be stopped immediately.

To avoid all the harmful effects of overloading and unbalancing of load, the load on a diesel engine should be equally distributed among the working cylinders; and no cylinder, or the engine itself, should ever be overloaded. In conclusion, the correct procedure to follow in balancing an engine is:

1. Maintain the engine in proper mechanical condition.

3. Operate the engine within the temperature limits specified in the instructions.

5. Train yourself to detect a bad condition by the senses of touch and hearing.

All rotating parts are subjected to two kinds of unbalance. They are called static unbalance and dynamic unbalance. The unbalanced condition in both cases can be readily determined and corrected by counterbalancing.

A static balancing test is conducted by placing the two ends of the rotating part on perfectly smooth, horizontal, and parallel rails. If statically unbalanced, the part will roll on the rails until its center of gravity reaches its lowest position and then it will come to rest. If, however, its center of gravity lies along its axis it will remain at rest when placed in any position, and it is then in static balance.

It frequently occurs that the center of gravity of a body lies in its axis of rotation but that its irregular shape or composition generates a disturbing force when the body is rotated. In this case the body would be in static balance and in dynamic unbalance. In general, before balancing, most rotating parts are in both static and dynamic unbalance.

In all cases, complete balancing can be obtained by attaching weights to the rotating body, if the position and degree of unbalancing are known. For determining this unbalance all naval shipyards are equipped with balancing machines. Experience with large and high-speed machinery has shown that balancing machines show good results but do not insure against

In any event, all rotating parts of the engine should be as accurately balanced as possible.

9D2. Flywheels. A flywheel stores up energy, the amount of which depends upon the rotating speed, the weight, and the diameter of the wheel. In most marine engines heavy flywheels are not necessary, as the other rotating masses on the shaft serve the same purpose. These masses are the clutch and generator, and with a large number of cylinders firing, the power stroke is smoother, and there is less need for a flywheel.

The flywheel serves three purposes, namely:

1. To prevent the engine from stalling when running at idling speed.

2. To reduce the variations in speed at all loads.

3. To help carry the engine over centers when starting.

When the speed of the shaft tends to increase, the flywheel absorbs energy. When it tends to decrease, the flywheel gives up its energy to the shaft in an effort to keep it rotating at a uniform speed.

9D3. Torsional vibrations. The twisting and untwisting of the shaft system result in torsional vibrations. All shafts have some flexibility and with weights attached to them, such as pistons, gears and camshafts in diesel engines, they have what is known as a natural fixed frequency. When the frequency of the power stroke impulses coincides with the natural frequency of the entire shaft system, a torsional vibration is produced, and the shaft is then said to be

Without going into further detail, it is sufficient to say that torsional critical speeds depend upon the number of power impulses per revolution and the natural rate of vibration of the combined shaft system. Special instruments are available for determining the degree of torsional vibration and the natural frequency of any particular shaft system.

To change the range or point of maximum vibration of the critical speeds for a given installation it is necessary to make a change in the masses on the elastic shaft system. It is evident therefore that in engines operating at a constant speed, it is much simpler to change the natural frequency in order to avoid dangerous critical speeds than it is in a marine engine requiring a wide range of operating speeds.

Critical speeds and mode of vibrations are determined with the aid of an instrument recording torsional vibrations. The engine builders calculate the critical speeds and furnish a guarantee that, with the engine coupled to the load for which it is designed, no dangerous critical speeds will occur within the operating speeds.

Torsional vibrations need not necessarily shake the framing of the engine and may not even be noticeable to the operator. This fact has been borne out in several casualties in which the crankshaft broke without warning. Excessive wear of gears or of attached auxiliaries and repeated breakage of shafting or other parts attached to it can very well be caused by torsional vibrations. Most installations in naval vessels have been checked and tested to determine the exact location of torsional vibrations, their amplitudes, and frequencies.

9D4. Flexural vibrations. The bending of the parts of the engine framing such as the

9D5. Torsional vibration dampening. There are certain forces acting in resistance to torsional vibrations. These forces are due to the friction of the bearings that carry the shafting and the work absorbed in the metal of the shaft in resisting the twisting called hysteresis. Propellers in the water are the most influential factor. All of these forces may be said to be the result of natural causes, and they act to dampen out, or reduce the amplitude of the torsional vibrations. In addition to these natural forces, there are other methods employed to reduce or eliminate the severity of the vibrations. This may be accomplished by changing the firing order of the cylinders in the engine, or by changing the rotating weights, or the flexibility of the shaftings.

In addition to the above dampening factors and methods there are various types of commercial torsional vibration dampeners, such as that used on the F-M 38D 8 1/8 10-cylinder engine. Each such dampener must be designed for a specific shaft system operating with a particular type engine. Vibration dampeners are usually located at or near a point of maximum torsional vibration amplitude along the shaft, generally at the forward end of an engine.

There are several different types of dampeners. All, however, accomplish the same purpose. They tend to reduce the swinging motion of the shaft. This is accomplished by having a freely rotating disk or disks acting against a fixed disk which creates friction and thereby acts as a brake. This prevents the shaft, from twisting and untwisting while rotating on its axis.

9E2. Types of engine indicators. There are two general classes of engine pressure recording indicators. In the first, the instrument measures graphically the cylinder pressure and at the same time indicates the position of the piston at any point of its stroke or cycle. In other words, the indicator draws a diagram of the pressure in the cylinder with respect to the movement of the piston. Since the movement of the piston is a measure of the volume displaced, the diagram is drawn to the ordinates of pressure and volume. In the second general class, the indicator records the maximum pressures only.

Figure 9-5 shows the fundamental principle of the operation of an engine indicator in which the movement of the piston is recorded. The indicator equipment includes a small cylinder that can be attached to the main working cylinder of the engine, a piston and rod that work in this small cylinder, with a pencil on the end of the rod. The pencil point bears on the paper tacked to the drum which is moved by hook and string over a pulley. Any pressure in the working cylinder enters the indicator cylinder and forces the small indicator piston and pencil in a vertical direction at the same time the main piston moves the card in a horizontal direction by means of the string and pulley.

It is readily seen that any vertical distance on the diagram will represent pressure, and the

This type of indicator is little used by operating personnel on fleet type submarines today, mainly because there is no provision made on modern engines for the attachment of the equipment necessary to take the indicator card, and also because there are no means of compression adjustment other than complete overhaul of the engine.

The other type of indicator (indicating maximum pressures only) is used to some extent for taking maximum cylinder pressures, to check against manufacturer's test data and previous shipboard pressure tests. The two most commonly used indicators of this type are the Premax indicator and the Kiene indicator.

The method normally used to check the equal distribution of power among the various cylinders is to compare the exhaust gas temperatures of the cylinders by means of thermocouples placed in the exhaust elbows of each cylinder. Pyrometer readings have proved to be a good check on the general running conditions of an engine, and the records of exhaust gas temperatures are of great value in conjunction with indicator readings as aids in getting the best results from a diesel engine. However, even though the exhaust temperatures are normal, the engine at times may not develop its rated horsepower.

9E3. Premax indicator. The Premax indicator is an instrument for determining cylinder

The cylinder discharge passes through the indicator plug up through the filler, screen, and seat piece. This raises the valve, allowing the gas to pass through the drilled holes in the guide piece into the pressure chamber and on to the gage. The action of the gas in the curved tube of the gage tends to straighten the tube, thereby moving the gage needle and recording the pressure on a calibrated scale.