8.1.1 Advantages of diesel power plant over steam power plant

8.1.2 Disadvantages of diesel power plant over steam power plant

8.1.3 Mean effective pressure (M.E.P.)

In a reciprocating engine it is difficult to isolate the positive and negative pressure in the cycle, since the processes are carried out in a single component. Hence, while comparing air-standard cycles of reciprocating engines a term mean effective pressure is usually referred to in lieu of the work ratio. Mean Effective Pressure, or in short M.E.P., is defined as the mean pressure that, if imposed on the pistons uniformly from the top to the bottom of each power stroke, would produce the net work of the cycle. A cycle operating with large M.E.P. will produce large work output per unit of swept volume. As a result size of an engine based on this cycle will be small for a given output.

For naturally aspirated four-stroke diesel engines, the maximum M.E.P. is in the range of 700 to 900 kPa. Two-stroke diesel engines have comparable values of M.E.P., while very large low speed diesel engines may have a M.E.P. up to 1900 kPa.

8.1.4 Four-stroke engine

A four-stroke engine comprises of the following:

8.1.5 Two-stroke engine

In a two-stroke engine the induction and exhaust strokes are eliminated and part of the compression and expansion strokes are used for exhaust and induction. At the end of the compression stroke the inlet port opens and air is admitted to the cylinder, while at the end of the expansion stroke the exhaust port is opened to allow blow-down of gases. Eventually the piston stroke of this engine is longer than that of a four-stroke engine of the same capacity.

The relative advantages/disadvantages of four-stroke and two-stroke engines are as follows:

8.1.6 Advantages of four-stroke engines

8.1.7 Advantages of two-stroke engine

Two-stroke engines have one major disadvantage – leakage of scavenge air through the exhaust ports accounts for a loss of air almost equal to 50% of the swept volume of the cylinder.

8.2.1 Diesel cycle

Figure 8.2 and Figure 8.3 depict the P-v and T-s diagrams for an ideal Diesel cycle, where P is the pressure and v is the specific volume. The ideal Diesel cycle follows four distinct processes: compression, combustion, expansion, and cooling. The cycle follows the numbers 1–4 in the clockwise direction.

The stages of these four processes are discussed in the following. Note that theoretically each of the strokes of the cycle complete at either top dead center (TDC) or bottom dead center (BDC). However, due to delay in the opening and closing of the inlet and exhaust valves, the inertia of inlet air and exhaust gas, each of the strokes invariably begin and end beyond the TDC and BDC.

8.2.1.1 Compression

The cycle is started by drawing air at ambient conditions into the engine (Figure 8.4). This air is then compressed adiabatically by moving the piston upward in the cylinder from BDC to TDC (Figure 8.5). This compression raises the temperature of the air to a level where the fuel mixture, which is formed by injecting fuel once the air is compressed, spontaneously ignites. It is in this part of the cycle that work is applied to the air. This compression is considered to be isentropic.

The ratio of the volume of the working fluid inducted into the cylinder before the compression process (v₁) to its volume after the compression (v₂) is defined as the volumetric compression ratio, r_c (=v₁/v₂).

8.2.1.2 Combustion (Figure 8.6)

Next, the fuel is injected into the combustion chamber during the final stage of compression, thereby adding heat to the air by combustion of fuel. This process begins just as the piston leaves its TDC position. Because the piston is moving during this part of the cycle, the heat addition process is called isochoric (constant volume). The combustion process, however, takes place at constant pressure.

8.2.1.3 Expansion (Figure 8.7)

In the Diesel cycle, fuel is burned to heat compressed air and as the rapidly burning mixture attempts to expand within the cylinder, it generates a high pressure that forces the piston downward from the TDC position to the BDC position in the cylinder. It is in this phase that the cycle contributes its useful work, rotating the engine crankshaft. This stage of an ideal Diesel cycle is considered as isentropic.

The ratio of the volume of the working fluid completing the expansion process (v₄=v₁) to its volume at the beginning of expansion (v₃) is called the volumetric expansion ratio, r_e (=v₁/v₃).

The expanded mixture or gas is then ejected from the cylinder to the environment, which is replaced with a fresh supply of cold air from ambient conditions. Since this happens when the piston is at the TDC position in the cycle and is not moving, this process is isochoric (no change in volume). During actual operation there is a point in the cycle when both inlet and exhaust valves remain open simultaneously.

8.2.1.4 Thermal efficiency [5]

As explained above the compression and power strokes of an ideal Diesel cycle are adiabatic, hence the thermal efficiency of this ideal cycle can be calculated from the constant-pressure and constant-volume processes. Accordingly, the heat input (Q_in), heat output (Q_out), and thermal efficiency (η_th) can be calculated if the temperatures and specific heats of various end states are known. Assuming constant specific heats for the air and considering unit mass flow of the fluid, using Figure 8.3 we can write:

$Q_{in}=C_p(T_3-T_2)$ (8.1)

(8.1)

$Q_{out}=C_v(T_1-T_4)$ (8.2)

(8.2)

${\eta }_{th}=\frac{Q_{in}+Q_{out}}{Q_{in}}$ (8.3)

(8.3)

or,

${\eta }_{th}=1+\frac{Q_{out}}{Q_{in}}=1+\frac{C_v(T_1-T_4)}{C_p(T_3-T_2)}$ (8.4)

(8.4)

where

C_p=Specific heat at constant pressure

C_v=Specific heat at constant volume

T=Temperature at various end states of the cycle

From Figure 8.2 it can also be written that the compression ratio r_c=v₁/v₂ and the expansion ratio r_e=v₁/v₃. Now, using the ideal gas law Pv=nRT and γ=C_p/C_v and for adiabatic condition Pv^γ=constant, Eq. 8.4 can be written as

${\eta }_{th}=1-\frac{1}{\gamma }\frac{(T_4-T_1)}{(T_3-T_2)}$ (8.5)

(8.5)

For adiabatic compression

$\frac{T_2}{T_1}={r_c}^{(\gamma -1)}$ (8.6)

(8.6)

therefore

$T_1=\frac{T_2}{{r_c}^{(\gamma -1)}}$ (8.7)

(8.7)

For constant-pressure process

$\frac{T_3}{T_2}=\frac{v_3}{v_2}=\frac{v_3}{v_1}*\frac{v_1}{v_2}=\frac{r_c}{r_e}$ (8.8)

(8.8)

hence,

$T_3=T_2\frac{r_c}{r_e}=T_1\frac{r_c^{\gamma }}{r_e}$ (8.9)

(8.9)

Again, for adiabatic expansion,

$\frac{T_3}{T_4}={r_e}^{(\gamma -1)}$ (8.10)

(8.10)

or,

$T_4=\frac{T_3}{r_e^{(\gamma -1)}}=T_2\frac{r_c}{r_e}\frac{1}{r_e^{(\gamma -1)}}=T_2\frac{r_c}{r_e^{\gamma }}=T_1{\left(\frac{r_c}{r_e}\right)}^{\gamma }$ (8.11)

(8.11)

Substituting the values of T₁, T₃ and T₄ from Eq. 8.7, Eq. 8.9, and Eq. 8.11, respectively, into Eq. 8.5 we arrive at

${\eta }_{th}=1-\frac{1}{\gamma }\frac{\left(T_2\frac{r_c}{r_e^{\gamma }}-\frac{T_2}{r_c^{(\gamma -1)}}\right)}{\left(T_2\frac{r_c}{r_e}-T_2\right)}=1-\frac{1}{\gamma }\frac{\left(\frac{r_c}{r_e^{\gamma }}-\frac{1}{r_c^{(\gamma -1)}}\right)}{\left(\frac{r_c}{r_e}-1\right)}=1-\frac{1}{\gamma }\frac{\left(\frac{1}{r_e^{\gamma }}-\frac{1}{r_c^{\gamma }}\right)}{\left(\frac{1}{r_e}-\frac{1}{r_c}\right)}$

or,

${\eta }_{th}=1-\frac{1}{\gamma }\frac{(r_e^{-\gamma }-r_c^{-\gamma })}{(r_e^{-1}-r_c^{-1})}$ (8.12)

(8.12)

Sometimes a “cut-off” ratio, as follows, is used to carry out various calculations:

$r_v=\frac{v_3}{v_2}=\frac{v_1}{v_2}*\frac{v_3}{v_1}=\frac{r_c}{r_e}$ (8.13)

(8.13)

Inputting the value of Eq. 8.13 into Eq. 8.12

${\eta }_{th}=1-\frac{1}{\gamma }\frac{r_c^{-\gamma }\left\{{\left(\frac{r_e}{r_c}\right)}^{-\gamma }-1\right\}}{r_c^{-1}\left\{{\left(\frac{r_e}{r_c}\right)}^{-1}-1\right\}}=1-\frac{r_c^{(1-\gamma )}}{\gamma }\frac{\left\{{\left(\frac{1}{r_v}\right)}^{-\gamma }-1\right\}}{{\left(\frac{1}{r_v}\right)}^{-1}-1}=1-\frac{r_c^{(1-\gamma )}}{\gamma }\frac{(r_v^{\gamma }-1)}{(r_v-1)}$ (8.14)

(8.14)

Equations 8.12 and 8.14 give the ideal thermal efficiency of the Diesel cycle, which depends on the compression ratio as well as either the expansion ratio or the cut-off ratio, but not on the peak temperature T₃. However, the thermal efficiency of actual Diesel cycle will be much lower than the ideal efficiency due to heat and friction losses, power absorbed by auxiliaries, e.g., lubricating oil pumps, cooling water pumps, fuel pumps, radiator fans, etc.

8.2.1.5 Mean effective pressure (M.E.P.)

As explained in Section 8.1.3 that M.E.P. is defined as the mean pressure, which, if imposed on the pistons uniformly from the top to the bottom of each power stroke, would produce the net work of the cycle.

In a cycle net work done=heat added – heat rejected.

Therefore,

$W=C_p(T_3-T_2)-C_v(T_4-T_1)$ (8.15)

(8.15)

Again, a change in volume,

$v_1-v_2=v_1(1-\frac{v_2}{v_1})=\frac{R{T}_1}{P_1}(1-\frac{1}{r_c})=\frac{(C_p-C_v)T_1}{P_1}\left(\frac{r_c-1}{r_c}\right)=\frac{C_v(\gamma -1)T_1}{P_1}(\frac{r_c-1}{r_c})$ (8.16)

(8.16)

Using the values from Eq. 8.15 and Eq. 8.16,

$M.E.P.=\frac{W}{v_1-v_2}=\frac{C_p(T_3-T_2)-C_v(T_4-T_1)}{\frac{C_v(\gamma -1)T_1}{P_1}\left(\frac{r_c-1}{r_c}\right)}=\frac{P_1{r}_c}{(r_c-1)}\frac{1}{(\gamma -1)}\left\{\gamma \frac{(T_3-T_2)}{T_1}-\frac{(T_4-T_1)}{T_1}\right\}$ (8.17)

(8.17)

Inputting the values from Eq. 8.7, Eq. 8.9, and Eq. 8.11 into Eq. 8.17

$M.E.P.=\frac{P_1{r}_c}{(r_c-1)}\frac{1}{(\gamma -1)}\left\{\gamma \left(\frac{r_c^{\gamma }}{r_e}-r_c^{(\gamma -1)}\right)-{\left(\frac{r_c}{r_e}\right)}^{\gamma }+1\right\}$

or,

$M.E.P.=\frac{P_1{r}_c}{(r_c-1)(\gamma -1)}\{\gamma r_c^{(\gamma -1)}(r_v-1)-(r_v^{\gamma }-1)\}$ (8.18)

(8.18)

Example 8.1

1 kg of air at 298 K temperature and 101.3 kPa pressure is taken through the ideal Diesel cycle of Figure 8.1. The compression ratio, r_c, of the cycle is 15 and the heat added is 1900 kJ. Find the ideal thermal efficiency and M.E.P. of this cycle.

(Assume R=0.287 kJ/(kg.K), C_p=1.005 kJ/kg/K, and γ=1.4 for air and conversion; factors: 1 kJ=102 m.kg and 1 kPa=102 kg/m^2.)

Solution: For the unit mass flow the initial volume of the air is

$v_1=\frac{R{T}_1}{P_1}=\frac{0.287*298}{101.3}=0.844m^3$

After compression

$T_2=T_1*{r_c}^{(\gamma -1)}=298*{15}^{0.4}=880.34K$

and

$v_2=\frac{0.844}{15}=0.056m^3$

With heat added

$Q_{in}=C_p(T_3-T_2)$

or,

$1900=1.005(T_3-880.34)$

hence,

$T_3=2770.89K$

The expansion ratio is

$r_e=\frac{v_4}{v_3}=\frac{v_1}{v_3}=\frac{v_1}{v_2}*\frac{v_2}{v_3}=\frac{v_1}{v_2}*\frac{T_2}{T_3}=15*\frac{880.34}{2770.89}=4.766$

hence,

$T_4=T_3\frac{1}{{r_e}^{(\gamma -1)}}=2770.89\frac{1}{{4.766}^{0.4}}=1483.74K$

With heat rejected

$Q_{out}=C_v(T_1-T{}_4)=\frac{1.005}{1.4}(298.00-1483.74)=-851.2kJ/kg$

Net work

$W=Q_{in}+Q_{out}=1900-851.2=1048.8kJ/kg$

Therefore, the ideal thermal efficiency is

${\eta }_{th}=\frac{1048.8}{1900.0}*100=55.2\%$

and

$M.E.P.=\frac{1048.8}{(0.844-0.056)}=1330.96kPa=1.331MPa$

An alternative solution of η_th using Eq. 8.12:

${\eta }_{th}=1-\frac{1}{1.4}\frac{({4.766}^{-1.4}-{15}^{-1.4})}{({4.766}^{-1}-{15}^{-1})}=1-\frac{1}{1.4}\frac{(0.1124-0.0226)}{(0.2098-0.0667)}=1-\frac{0.0898}{0.2003}=0.5518,i.e.55.18\%$

Example 8.2

Considering the unit mass flow of the fluid determine the thermal efficiency, the final temperature of fluid after expansion, and the compression ratio of an air-standard Diesel cycle, in which the inlet pressure and temperature are 100 kPa and 288 K, respectively. The heat addition to the cycle is 1045.9 kJ and the final temperature after heat addition is 2073 K. (Given C_p=1.005 kJ/kg/K, C_v=0.718 kJ/kg/K, and γ=1.4 for air.)

Solution: From Eq. 8.1, 1045.9=1.005(2073 – T₂)

Hence, T₂=1032.3 K

From Eq. 8.6, r_c^0.4=1032.3/288

Wherefrom, r_c=24.32

From Eq. 8.8 and Eq. 8.13, the cut-off ratio is

$r_v=2073/1032.3=2.008$

From Eq. 8.11

$T_4=288*{2.008}^{1.4}=764.3K$

Thus, from Eq. 8.14, the thermal efficiency of the air-standard Diesel cycle is

${\eta }_{th}=\left\{1-\frac{{24.32}^{-0.4}}{1.4}\left(\frac{{2.008}^{1.4}-1}{2.008-1}\right)\right\}*100=\left(1-\frac{1}{5.018}\frac{1.654}{1.008}\right)*100=67.3\%$

8.2.2 Otto cycle

The Otto cycle was developed by German engineer Nikolaus Otto in 1876. This cycle is very similar to the Diesel cycle, except that the Otto cycle forms the basis of a spark-ignition (S.I.) engine, which requires a spark to begin combustion of fuel. Figure 8.8 represents a P-V diagram of an Otto cycle.

As in an ideal Diesel cycle four distinct processes are also followed in the Otto cycle: reversible adiabatic compression, addition of heat at constant volume, reversible adiabatic expansion, and rejection of heat at constant volume as follows:

Stages of above four processes are described below:

8.2.2.1 Compression

In this part of the cycle as the piston moves from the TDC to the BDC, the inlet valve is open and the exhaust valve is closed. A fresh charge of fuel-air mixture is drawn into the cylinder (Figure 8.9). As the inlet valve closes the piston moves from the BDC upward to the TDC, and in the process compresses the fuel-air mixture (Figure 8.10). In this stroke work is contributed to the air. In an ideal Otto cycle, this compression is considered to be isentropic.

8.2.2.2 Combustion (Figure 8.11)

In this stage heat is added to the fuel-air mixture by combustion of fuel when the piston is at the TDC position. Combustion is initiated by a spark with the help of a spark plug, which is the most important and fundamental difference between the Otto cycle and the Diesel cycle. The piston during heat addition remains at a standstill and hence the process of heat addition is known as isochoric (constant volume).

8.2.2.3 Expansion (Figure 8.12)

While tending to expand, the rapidly burning fuel-air mixture generates high pressure thus forcing the piston to travel downward from the TDC to the BDC in the cylinder. It is in this stage that the cycle contributes its useful work, rotating the engine crankshaft. In an ideal Otto cycle this process is isentropic.

In an engine, the cooling process involves exhausting the gas from the engine to the environment and replacing it with fresh air. Since this happens when the piston is at the TDC position in the cycle and remains immobile, this process is also isochoric (no change in volume).

8.2.2.4 Thermal efficiency [5]

The thermal efficiency of the Otto cycle can be expressed by dividing external work by heat input to the working fluid as follows:

${\eta }_{th}=\frac{Q_{in}+Q_{out}}{Q_{in}}=1+\frac{Q_{out}}{Q_{in}}$ (8.19)

(8.19)

Hence, considering the unit mass flow of the fluid and assuming constant specific heats for the air, for constant volume heating and cooling we can use

${\eta }_{th}=1+\frac{C_v(T_1-T_4)}{C_v(T_3-T_2)}=1-\frac{T_4-T_1}{T_3-T_2}$ (8.20)

(8.20)

We know that for ideal gas Pv=nRT and γ=C_p/C_v and for adiabatic condition Pv^γ=constant. Using Figure 8.8 v₃=v₂ and v₄=v₁. Hence,

$\frac{T_2}{T_1}={\left(\frac{P_2}{P_1}\right)}^{\frac{\gamma -1}{\gamma }}={\left(\frac{v_1}{v_2}\right)}^{(\gamma -1)}$ (8.21)

(8.21)

or,

$T_1=\frac{T_2}{{\left(\frac{v_1}{v_2}\right)}^{(\gamma -1)}}$ (8.22)

(8.22)

Again,

$\frac{T_4}{T_3}={\left(\frac{P_4}{P_3}\right)}^{\frac{\gamma -1}{\gamma }}={\left(\frac{v_3}{v_4}\right)}^{(\gamma -1)}$ (8.23)

(8.23)

or,

$T_4=T_3{\left(\frac{v_2}{v_1}\right)}^{(\gamma -1)}=\frac{T_3}{{\left(\frac{v_1}{v_2}\right)}^{(\gamma -1)}}$ (8.24)

(8.24)

Using the values of Eq. 8.22 and Eq. 8.24 in Eq. 8.20 and expressing the compression ratio as r_c=v₁/v₂ the thermal efficiency of the Otto cycle changes to

${\eta }_{th}=1-\frac{1}{{\left(\frac{v_1}{v_2}\right)}^{(\gamma -1)}}=1-\frac{1}{r_c^{(\gamma -1)}}$ (8.25)

(8.25)

where

η_th=Thermal efficiency

Q_in=Heat added to the working fluid

Q_out=Heat rejected from the working fluid

C_p=Specific heat at constant pressure

C_v=Specific heat at constant volume

Equation 8.25 shows that the thermal efficiency of the Otto cycle depends only on the compression ratio and not on the peak temperature T₃.

Comparing Eqs. 8.12/8.14 and Eq. 8.25 it is noted that the formula for the thermal efficiency of the Diesel cycle is more complex than that of the Otto cycle. This is due to the fact that the heat addition in the Diesel cycle is at constant pressure and the heat rejection is at constant volume, while both heat addition and heat rejection in the Otto cycle take place at constant volume. For the same compression ratio the thermal efficiency of the Otto cycle is always greater than that of the Diesel cycle. In practice, however, the Diesel cycle yields higher thermal efficiency because it adopts higher compression ratios than those adopted by the Otto cycle, which is evident from Example 8.3.

Example 8.3

Determine the thermal efficiency, the final temperature after expansion, and the compression ratio of an air-standard Otto cycle for the same problem as given for Example 8.2.

Solution: From Eq. 8.19 and Eq. 8.20, 1045.9=0.718(2073 – T₂)

Hence, T₂=616.3 K

From Eq. 8.22, $r_c^{0.4}$=616.3/288=2.14

Therefore, r_c=6.7

From Eq. 8.24, T₄=2073/6.7^0.4=968.7 K

From Eq. 8.25 ${\eta }_{th}=\left(1-\frac{1}{{6.7}^{0.4}}\right)*100=53.3\%$

Comparing the results of Example 8.2 and Example 8.3 it is clear that the thermal efficiency of the air-standard Diesel cycle (67.3%) is much higher than that of the Otto cycle. This is because the compression ratio of the Diesel cycle (24.3) greatly exceeds the compression ratio of the Otto-cycle, which is 6.7.

In practice many reciprocating engines adopt dual or mixed cycles as shown in Figure 8.13, where heat is added in part in the constant volume process and the rest in the constant-pressure process.

To find out the thermal efficiency of the dual cycle, as before, let us assume constant specific heats of air and consider unit mass flow of the fluid. If the compression ratio between point 2 and point 3 is expressed as r_p=P₃/P₂ and as before the compression ratio between point 1 and point 2 as r_c=v₁/v₂ and the cut-off ratio between point 4 and point 3 as r_v=v₄/v₃, then the efficiency of this cycle could be expressed as follows:

${\eta }_{th}=1+\frac{Q_{out}}{Q_{in}}=1+\frac{C_v(T_1-T_5)}{C_p(T_4-T_3)+C_v(T_3-T_2)}=1+\frac{T_1-T_5}{\gamma (T_4-T_3)+(T_3-T_2)}$ (8.26)

(8.26)

Converting Eq. 8.26 from temperature to compression-ratio, pressure-ratio, and cut-off ratio

${\eta }_{th}=1-\frac{1}{r_c^{(\gamma -1)}}\frac{r_p{r}_v^{\gamma }-1}{(r_p-1)+\gamma r_p(r_v-1)}$ (8.27)

(8.27)

8.4.1 Fuel injection equipment

Correct functioning of the fuel injection equipment begets efficient combustion in a Diesel engine. Fuel is injected through nozzles at adequate pressure for proper atomization of fuel. The nozzle area should be small enough to allow fuel intake for complete vaporization without remaining in suspension as droplets. Depending on the size of the engine the atomizing pressure developed by the fuel injection pump/s may vary from 40 MPa to 100 MPa. The atomized fuel thus obtained mixes with pressurized air in the correct proportion in the combustion chamber so that complete combustion is ensured. This equipment ensures supply of metered quantity of fuel. The fuel injection equipment also controls fuel-air mixing rate according to the required power output. This equipment needs to be designed with high standards to limit gaseous emissions, noise, and smoke in compliance with the local environmental regulations.

8.4.2 Piston

In all internal combustion engines the design of a piston is very critical since it is the most vulnerable member. It is subjected to two prone attacks: one from the mechanical loading resulting from the maximum cylinder pressure and another thermal loading due to the heat of compression. Gas loads are transmitted to the crank shaft with the help of the piston. The design of a piston is influenced by the following parameters:

The sealing of gas and oil in a cylinder is provided by piston rings.

8.4.3 Governor

Like any rotary power-generating device every engine of a Diesel-generating plant is invariably provided with a governor to maintain the speed of the engine automatically by regulating the flow of fuel to meet the desired power output. A governor may be a hydro-mechanical or electronic type.

8.4.4 Starting system

To start a Diesel engine from standstill sufficient torque to the crankshaft has to be provided from the outside to rotate the engine at the desired speed, which is essentially done by a starting gear. The most common starting gear is the manual type, e.g., hand cranking or a rope and pulley system. It is simple, does not require an external source to break the starting torque, and its costs are low. The use of manual starting gear, however, is restricted to low-speed, small engines. The most commonly used starting gears for medium-speed and high-speed engines are electric. They consist of DC motors and battery packs to supply electric energy. In medium-speed and high-speed engines another starting gear that is widely used is pneumatic. An air motor or a power cylinder is supplied with pressurized air from an air compressor, which facilitates breaking the starting torque.

8.4.5 Intake silencer

When intake air from atmosphere rushes through the intake valve with high velocity it produces a hissing noise, which is intolerable to human ears. This noise is reduced by providing a silencer at the intake manifold.

8.4.6 Air filters

A filter is installed at the air intake to prevent wear of pistons, piston rings, cylinder, valves, etc., by airborne suspended impurities. An air filter also assists in the attenuation of intake noise.

8.4.7 Exhaust mufflers

When leaving the exhaust pipe exhaust gases produce an unbearable howling noise, which is much louder sound than intake noise. Hence, it is imperative to provide different means, e.g., expansion of exhaust gas, or change of direction of flow of gas, or cooling of gas with water, etc., at the exhaust system – for abatement of howling noise.

8.4.8 Supercharger

Supercharging of an engine can be defined as the artificial raising of the intake pressure of air above atmospheric pressure. The pressure is raised by the use of a compressor. If the compressor is driven by a turbine that operates on the exhaust gases from the engine it is called a turbocharger (Figure 8.14).

The objectives of supercharging are to:

8.5.1 Fuel oil system (Figure 8.15)

The fuel oil used in a Diesel engine should be completely free from water and mechanical impurities. Solid and liquid contaminants are cleaned before use to prevent damage to fuel pumps and engine components. A diesel engine is capable of burning a wide range of low-quality liquid. It should be ensured that heating of fuel to the required viscosity is carried out without developing thermal cracking. For start-up of a diesel engine either high-speed diesel (HSD) or light diesel oil (LDO) is used. For normal running, a heavier grade fuel oil, e.g., heavy fuel oil (HFO) or low sulphur heavy stock (LSHS), is generally used.

The fuel oil system deals with oil receipt, storage, treatment, and forwarding of oil to the Diesel engine. Fuel oil may be delivered to a plant by road tankers or railway wagons or if the plant is located near a port through the pipeline from the port. Oil delivered by road tankers or railway wagons is unloaded with the help of unloading pumps equipped with suction strainers. Connection is made with the tankers or wagons through screwed hoses. Pressurized fuel oil from unloading pumps is supplied through the pipeline to bulk fuel oil storage tanks for storing. These storage tanks are usually vertical, cylindrical, cone-roof steel tanks. While storing heavy-grade fuel oil, the minimum temperature in bulk oil storage tanks should be maintained at around 353 K by electric or steam heating to ensure the fluidity of the oil. However, for ensuring the required viscosity of heavy-grade fuel oil at the engine end the minimum temperature should be maintained in the treated oil storage tank at around 393 K.

The fuel system of large diesel engines generally includes the following:

The fuel feed pump supplies fuel from the treated oil storage tank/s and delivers it through a fine filter to the high-pressure fuel injection pumps. These pumps transfer oil to the injectors located on each cylinder.

8.5.2 Lubricating oil system (Figure 8.16)

The lubricating oil system of a Diesel engine typically includes an external lube oil circuit. The primary function of this circuit is to filter and cool the circulating oil. It also serves to heat, prime, and scavenge the oil when preparing the diesel engine for starting. An internal lube oil circuit supplies oil to individual units and parts of the diesel engine. The pipelines of internal circuit are inside the engine.

The external lube oil circuit in turn is typically divided into two groups, i.e., the main lube oil circuit and the auxiliary lube oil circuit. The main lube oil is either delivered by road tankers with the help of lube oil unloading pumps or downloaded from drums containing lube oil to the Diesel engine main lube oil sump. A lube oil transfer pump delivers lube oil from the lube oil sump through a metal-in-oil detector, oil filter, and oil cooler into a circulating oil tank. When the pressure in the system exceeds the maximum limit a by-pass valve diverts part of the lube oil directly to the circulating tank. The temperature of the lube oil leaving the oil cooler is controlled by adjusting the flow of cooling water through the oil cooler automatically by a thermo-regulator.

The temperature of the lube oil in the circulating oil tank is maintained with the help of an oil pump provided with oil heater, which cuts-in automatically in the event that the oil temperature drops down to 313 K and cuts-out when the temperature rises above 323 K.

A lube oil separator is used to clean the oil circulating in the lubrication system. Oil is drawn by the separator pump, either from the system or from the tank, and after separation it is returned to the circulating tank. A temperature of 328 K or more, as recommended, is maintained in this circuit, if required, with the help of an oil heater.

Oil in the internal lube oil circuit is circulated with the help of an engine-driven internal lube oil pump. This pump draws oil from the circulating tank and delivers it to the engine through an oil filter. Prior to starting the Diesel engine it is primed with oil by the priming pump. The internal circuit is exclusively used for “piston cooling and lubrication of crosshead and crankpin bearing” and “lubrication of camshaft main bearings.” Oil is drained from the pistons and bearings into the Diesel engine sump, from which the scavenge pump takes its suction and delivers oil back to the circulating tank. Separate lube oil pipelines connect to various lubricating points of major and minor accessories, drives, and the turbo-charger.

8.5.3 Cooling water system (Figure 8.17)

The Diesel engine cooling system generally consists of two circuits: the internal cooling circuit, also known as jacket cooling system, and the external cooling circuit. The water circulating along the internal circuit cools the Diesel engine, while the water of the external circuit serves to cool the oil, the water of the internal cooling circuit and the air for supercharging. Distilled or de-mineralized water treated with a corrosion inhibitor is generally used in the inner circuit, while either raw water or filtered water is adopted in the external circuit.

The inner circuit cooling water pump takes water from the Diesel engine sump and circulates water through the engine jacket, jacket water cooler, and the turbo-compressor. From the supply manifolds, water also flows to cool the exhaust manifold sections and the working cylinders. The water from the working cylinders is drained to return manifolds, from which it is directed to the pump suction line, completing the cycle. An expansion tank is connected to the suction side of the pump to ensure initial priming, as well as to ensure reliable functioning of the closed system when the water volume changes due to heating or leakage. An independent water-circulating pump is provided with an electric heater to warm up the engine with hot water, particularly when the engine remains idle. The heater automatically cuts-in when the engine water outlet temperature falls below about 303 K and cuts-out when the temperature exceeds 323 K.

The external cooling circuit is equipped with an outer circuit cooling water pump. This pump draws water from an evaporative mechanical-draft cooling tower sump and directs it through the lube oil cooler and internal cooling circuit water cooler and then returns to the cooling tower with the help of a water transfer pump. Cooling water from this circuit is supplied to air coolers, generator-bearing cooler, and compressor cooling system as well. The amount of water passing through the lube oil cooler is usually determined by a thermo-regulator control unit, to maintain the operating temperature of the oil flowing out of the cooler.

For small capacity engines, the inner circuit cooling water may be cooled by supplying air to a radiator with cooling air fans.

8.5.4 Compressed air system

The starting air of a diesel engine is delivered by the air compressor. The pressurized air is stored in vertically mounted air receivers. During start-up, pressurized air enters the Diesel engine main starting pipeline through the main starting valves to serve bank of cylinders. From the main starting valves air gets into the air distributors as well as into working cylinder-starting valves. Once the diesel engine starts operating on fuel, the starting valve is closed to cut the supply of air.

From the starting air receivers control air is led off through pressure-reducing valve to the oil and water thermo-regulators. This air then regulates the flow of water through the cooler or heat-exchanger by opening or closing the valves.

8.5.5 Firefighting system

In a Diesel power station fire may start accidentally due to short circuiting, ignition of fuel and lubricants or of common combustible materials, e.g., paper, cotton waste, etc. As soon as minor fires are detected portable/semi-portable fire extinguishers are used to extinguish it.

Oil fire is best tackled by a foam-type extinguisher along with a spray water system. The foam provides a blanket over the oil surface and cuts off the oxygen supply, which is essential for supporting combustion.

As an alternative means, sand stored in the bins near the dyke walls of fuel storage tanks can also be used to prevent outbreak of minor fire. Sand is sprinkled evenly on the spillage surfaces using shovels/buckets. Small fires can be handled successfully using sand. Ordinary water is not generally suitable for fighting oil fire, since oil has a tendency to float and spread further.

Electrical fire is fought with carbon dioxide fire extinguishers. Wherever practical, the electric supply to the equipment on fire should be switched off before using the extinguisher. In general, fire due to combustible materials such as cotton waste, paper, timber, etc., is tackled either by soda-acid type fire extinguishers or with water stored in fire buckets. In the event of a big fire, portable/semi-portable extinguishers may not extinguish the fire effectively. In such an event, the central firefighting team should be called as soon as possible.