**Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine**

Marek Dzida  *Gdansk University of Technology Poland* 

#### **1. Introduction**

44 Advances in Gas Turbine Technology

Wilke, I. & Kau, H.-P. (2004). A Numerical Investigation of the Flow Mechanisms in a High

Yamada, K.; Furukawa, M.; Nakano, T.; Inoue, M. & Funazaki, K. (2004). Unsteady Three-

Yi, W.; Huang, H. & Han, W. (2006). Design Optimization of Transonic Compressor Rotor

Vol. 126, No. 3, (July 2004), pp. 339-349, ISSN 0889-504X

53745

90155

pressure Compressor Front Stage With Axial Slots, *ASME Journal of Turbomachinery*,

Dimensional Flow Phenomena Due to Breakdown of Tip Leakage Vortex in a Transonic Axial Compressor Rotor, *Proceedings of ASME Turbo Expo 2004*, GT2004-

Using CFD and Genetic Algorithm, *Proceedings of ASME Turbo Expo 2006*, GT2006-

For years there has been, and still is, a tendency in the national economy to increase the efficiency of both the marine and inland propulsion systems. It is driven by economic motivations (rapid increase of fuel prices) and ecological aspects (the lower the fuel consumption, the lower the emission of noxious substances to the atmosphere). New design solutions are searched to increase the efficiency of the propulsion system via linking Diesel engines with other heat engines, such as gas and steam turbines. The combined systems implemented in marine propulsion systems in recent years are based mainly on gas and steam turbines (MAN, 2010). These systems can reach the efficiency exceeding 60% in inland applications. The first marine system of this type was applied on the passenger liner "Millenium". However, this is the only high-efficiency marine application of the combined propulsion system so far. Its disadvantage is that the system needs more expensive fuel, the marine Diesel oil, while the overwhelming majority of the merchant ships are driven by low-speed engines fed with relatively cheap heavy fuel oil. It seems that the above tendency will continue in the world's merchant navy for the next couple of years.

The compression-ignition engine (Diesel engine) is still most frequently used as the main engine in marine applications. It burns the cheapest heavy fuel oil and reveals the highest efficiency of all heat engines. The exhaust gas leaving the Diesel engine contains huge energy which can be utilised in another device (engine), thus increasing the efficiency of the entire system and reducing the emission of noxious substances to the atmosphere.

A possible solution here can be a system combined of a piston internal combustion engine and the gas and steam turbine circuit that utilises the heat contained in the exhaust gas from the Diesel engine. The leading engine in this system is the piston internal combustion engine. It seems that now, when fast container ships with transporting capacity of 8- 12 thousand TU are entering into service, the propulsion engines require very large power, exceeding 50-80 MW. On the other hand, increasing prices of fuel and restrictive ecological limits concerning the emission of NOx and CO2 to the atmosphere provoke the search for new solutions which will increase the efficiency of the propulsion and reduce the emission of gases to the atmosphere.

The ship main engines will be large low-speed piston engines that burn heavy fuel oil. At present, the efficiency of these engines nears 45 – 50%. For such a large power output

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

25

engine.

Marine Diesel Engine

30

35

40

**efficiency [%]**

45

50

55

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 47

50 55 60 65 70 75 80 85 90 95 100

Fig. 1. Part - load Efficiency of a Combined - Cycle Plant (GT&ST), Simple Gas Turbine and

Combined systems used in inland power blocks base on a gas turbine as the main unit and a steam turbine that utilises the steam produced in a waste heat boiler using the heat recovered from the gas turbine exhaust gas. All this provides opportunities for reaching high efficiency of the combined block. The exhaust gas leaving a marine low-speed Diesel engine contains smaller amount of heat, of an order of 30-40% of the energy delivered to the

Figure 1 shows sample efficiency curves of the combined gas turbine/steam turbine systems as functions of power plant load, compared to the gas turbine operating in a simple open

The efficiency curves in Fig. 1 show that the combined cycle gas turbine/steam turbine system has the highest efficiency for maximal loads (maximal efficiency levels for these circuits reach as much as 60%). Gas turbines operating in the simple open circuit have the lowest efficiency (average values of 33÷35%, and maximal values reaching 40%). Lowspeed Diesel engines have the efficiency of an order of 47÷50%. It is also noticeable that the Diesel engine curve is relatively flat. This is of special importance in case of marine propulsion systems which operate at heavily changing loads. For the combined cycle gas turbine/steam turbine systems and the gas turbines operating in the simple open circuit the relative efficiency decrease is equal to 15÷20% when the load decreases from 100% to 50%. For the low-speed Diesel engine these numbers are equal to 1÷2%. This property of the Diesel engine, along with the ability to utilise additional heat contained in its exhaust gas, makes the engine the most applicable in marine propulsion systems operating in heavily changing load conditions. The amount of heat contained in the

**\_\_\_\_\_** CC Plant with Variable Inlet Guide, **\_\_\_\_\_** CC Plant without Variable Inlet Guide

**\_\_\_\_\_** Simple Gas Turbine, **\_\_\_\_\_** Marine Diesel Engine

circuit and the marine low-speed Diesel engine.

**N/No [%]**

ranges, the exhaust gas leaving the engine contains huge amount of heat available for further utilisation.

The proposed combined system consisting of a piston internal combustion engine, a gas turbine and a steam turbine can also be used for engines of lower power, ranging between 400 ÷900 kW. For those power ranges a use of low-boiling media of organic-based refrigerant type instead of water (steam) in the steam cycle seems to be a reasonable solution. Piston internal combustion engines of this power range are used on coasting vessels, or in the inland water transport, for instance for driving cargo barges. On the other hand in inland applications the power blocks fired with solid, liquid, or gas fuels are in almost 100% the systems with steam or gas turbines.

In the Central Europe, Poland for instance, the basic fuel in power engineering is coal. Conventional electric power plants have the efficiency of an order of 38-42%, and emit large volumes of CO2 , NOx and/or SOx. In order to decrease the amount of noxious substances emitted to the atmosphere and reduce the cost of production of the electric energy, combined systems are in use - consisting of gas turbines with a steam turbine circuit.

On the other hand, the combined turbine power plants can be complemented by electric power plants with a Diesel engine as the main propulsion. The exhaust gas leaving the engine contains about 30-40% of the heat delivered to the engine in the fuel. Using the heat from the exhaust gas in the gas and steam turbine circuit will increase the efficiency of the entire combined system. For large powers of piston internal combustion engines, the additional gas and steam turbine circuit is a source of measurable economic savings in electric energy production. Moreover, in large-power piston internal combustion engines we can additionally use the low-temperature waste heat, for instance for heating the communal water (Dzida, 2009). In the seaside areas with no large electric power plants, a mobile power plant situated on a platform close to the coast reveals additional advantages:


#### **2. Concept of a combined system**

Combined propulsion systems are used in marine engineering mostly in fast specialpurpose ships and in the Navy, as the systems being a combination of a Diesel engine and gas turbines (CODAG, CODOG) or solely gas turbines (COGOG, COGAG). The propulsion system of the passenger liner "Millenium" uses a COGES-type system which improved the efficiency and operating abilities of the ship. The system consists of a gas turbine and a steam turbine which drive an electric current generator, while the propeller screws are driven by electric motors. In this system the steam turbine circuit is supplied with the steam generated in the waste heat boiler supplied with the exhaust gas from the gas turbines.

ranges, the exhaust gas leaving the engine contains huge amount of heat available for

The proposed combined system consisting of a piston internal combustion engine, a gas turbine and a steam turbine can also be used for engines of lower power, ranging between 400 ÷900 kW. For those power ranges a use of low-boiling media of organic-based refrigerant type instead of water (steam) in the steam cycle seems to be a reasonable solution. Piston internal combustion engines of this power range are used on coasting vessels, or in the inland water transport, for instance for driving cargo barges. On the other hand in inland applications the power blocks fired with solid, liquid, or gas fuels are in

In the Central Europe, Poland for instance, the basic fuel in power engineering is coal. Conventional electric power plants have the efficiency of an order of 38-42%, and emit large volumes of CO2 , NOx and/or SOx. In order to decrease the amount of noxious substances emitted to the atmosphere and reduce the cost of production of the electric energy,

On the other hand, the combined turbine power plants can be complemented by electric power plants with a Diesel engine as the main propulsion. The exhaust gas leaving the engine contains about 30-40% of the heat delivered to the engine in the fuel. Using the heat from the exhaust gas in the gas and steam turbine circuit will increase the efficiency of the entire combined system. For large powers of piston internal combustion engines, the additional gas and steam turbine circuit is a source of measurable economic savings in electric energy production. Moreover, in large-power piston internal combustion engines we can additionally use the low-temperature waste heat, for instance for heating the communal water (Dzida, 2009). In the seaside areas with no large electric power plants, a mobile power



Combined propulsion systems are used in marine engineering mostly in fast specialpurpose ships and in the Navy, as the systems being a combination of a Diesel engine and gas turbines (CODAG, CODOG) or solely gas turbines (COGOG, COGAG). The propulsion system of the passenger liner "Millenium" uses a COGES-type system which improved the efficiency and operating abilities of the ship. The system consists of a gas turbine and a steam turbine which drive an electric current generator, while the propeller screws are driven by electric motors. In this system the steam turbine circuit is supplied with the steam generated in the waste heat boiler supplied with the exhaust gas from the gas turbines.

combined systems are in use - consisting of gas turbines with a steam turbine circuit.

plant situated on a platform close to the coast reveals additional advantages: - increasing production of electric energy in the so-called distributed system,




further utilisation.

favour of liquid fuels,

water balance aspects,

**2. Concept of a combined system** 

almost 100% the systems with steam or gas turbines.

Fig. 1. Part - load Efficiency of a Combined - Cycle Plant (GT&ST), Simple Gas Turbine and Marine Diesel Engine

**\_\_\_\_\_** Simple Gas Turbine, **\_\_\_\_\_** Marine Diesel Engine

Combined systems used in inland power blocks base on a gas turbine as the main unit and a steam turbine that utilises the steam produced in a waste heat boiler using the heat recovered from the gas turbine exhaust gas. All this provides opportunities for reaching high efficiency of the combined block. The exhaust gas leaving a marine low-speed Diesel engine contains smaller amount of heat, of an order of 30-40% of the energy delivered to the engine.

Figure 1 shows sample efficiency curves of the combined gas turbine/steam turbine systems as functions of power plant load, compared to the gas turbine operating in a simple open circuit and the marine low-speed Diesel engine.

The efficiency curves in Fig. 1 show that the combined cycle gas turbine/steam turbine system has the highest efficiency for maximal loads (maximal efficiency levels for these circuits reach as much as 60%). Gas turbines operating in the simple open circuit have the lowest efficiency (average values of 33÷35%, and maximal values reaching 40%). Lowspeed Diesel engines have the efficiency of an order of 47÷50%. It is also noticeable that the Diesel engine curve is relatively flat. This is of special importance in case of marine propulsion systems which operate at heavily changing loads. For the combined cycle gas turbine/steam turbine systems and the gas turbines operating in the simple open circuit the relative efficiency decrease is equal to 15÷20% when the load decreases from 100% to 50%. For the low-speed Diesel engine these numbers are equal to 1÷2%. This property of the Diesel engine, along with the ability to utilise additional heat contained in its exhaust gas, makes the engine the most applicable in marine propulsion systems operating in heavily changing load conditions. The amount of heat contained in the

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

**\_\_\_\_\_\_** Gas Turbine **\_\_\_\_\_\_** Marine Diesel Engine

propeller screw, and for covering all-ship needs.

Fig. 4. Concept of the combined propulsion system

Fig. 3. Related exhaust gas mass flow rate as a function of power plant load

5

10

15

**Exhaust gas flow [kg/kWh]**

20

25

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 49

40 50 60 70 80 90 100 110

The exhaust gas from the turbocharger and the power turbine flows to the waste heat boiler installed in the main engine exhaust gas path, before the silencer. The waste heat boiler produces the steam used both for driving the steam turbine that passes its energy to the

**N/No [%]**

exhaust gas from the gas turbine is approximately equal to 60÷65%, i.e. more than in piston engines, which results from lower exit temperature and less intensive flow of the exit gas leaving the Diesel engine, Figs. 2 and 3.

Fig. 2. Temperatures of the exhaust gas from the Diesel engine and the gas turbine as function of power plant load

The exit temperatures of the exhaust gas from the gas turbines range between 450÷6000C, on average, while those from the low-speed Diesel engines are of an order of 220÷3000C. In the gas turbines, decreasing the load remarkably decreases the temperature of the exhaust gas, while in the Diesel engine these changes are much smaller, and the temperature initially decreases and then starts to increase for low loads.

This property of the steam turbine circuit in the combined system with the Diesel engine for partial loads makes it possible to keep the live steam temperature at a constant level within a wide range of load. The related exhaust gas mass flow rate mg/N [kg/kWh] changes only by about 5% in the Diesel engine when the load changes from 100% to 50%, while in the gas turbine this parameter changes by about 55% for the same load change, Fig. 3.

The combined propulsion system with the low-speed piston internal combustion engine used as the main engine and making use of the heat from the engine exhaust gas is shown in Fig. 4, (Dzida, 2009; Dzida & Mucharski, 2009; Dzida et al., 2009).

The exhaust gas flows leaving individual main engine cylinders are collected in the exhaust manifold and passed to the constant-pressure turbocharger. Due to high turbocharger efficiency ranges (MAN, 2010; Schrott, 1995), the scavenge air can be compressed using the energy contained only in part of the exhaust gas flow. The remaining part of the exhaust gas flow can be expanded in an additional gas turbine, the so-called power turbine, which additionally drives, via a gear, the propeller screw or the electric current generator.

exhaust gas from the gas turbine is approximately equal to 60÷65%, i.e. more than in piston engines, which results from lower exit temperature and less intensive flow of the

40 50 60 70 80 90 100 110

The exit temperatures of the exhaust gas from the gas turbines range between 450÷6000C, on average, while those from the low-speed Diesel engines are of an order of 220÷3000C. In the gas turbines, decreasing the load remarkably decreases the temperature of the exhaust gas, while in the Diesel engine these changes are much smaller, and the temperature initially

This property of the steam turbine circuit in the combined system with the Diesel engine for partial loads makes it possible to keep the live steam temperature at a constant level within a wide range of load. The related exhaust gas mass flow rate mg/N [kg/kWh] changes only by about 5% in the Diesel engine when the load changes from 100% to 50%, while in the gas

The combined propulsion system with the low-speed piston internal combustion engine used as the main engine and making use of the heat from the engine exhaust gas is shown in

The exhaust gas flows leaving individual main engine cylinders are collected in the exhaust manifold and passed to the constant-pressure turbocharger. Due to high turbocharger efficiency ranges (MAN, 2010; Schrott, 1995), the scavenge air can be compressed using the energy contained only in part of the exhaust gas flow. The remaining part of the exhaust gas flow can be expanded in an additional gas turbine, the so-called power turbine, which

additionally drives, via a gear, the propeller screw or the electric current generator.

turbine this parameter changes by about 55% for the same load change, Fig. 3.

Fig. 4, (Dzida, 2009; Dzida & Mucharski, 2009; Dzida et al., 2009).

Fig. 2. Temperatures of the exhaust gas from the Diesel engine and the gas turbine as

**N/No [%]**

exit gas leaving the Diesel engine, Figs. 2 and 3.

\_\_\_\_\_ Gas Turbine \_\_\_\_\_ Marine Diesel Engine

decreases and then starts to increase for low loads.

function of power plant load

250

300

350

400

**Temperature [ oC ]**

450

500

550

Fig. 3. Related exhaust gas mass flow rate as a function of power plant load

The exhaust gas from the turbocharger and the power turbine flows to the waste heat boiler installed in the main engine exhaust gas path, before the silencer. The waste heat boiler produces the steam used both for driving the steam turbine that passes its energy to the propeller screw, and for covering all-ship needs.

Fig. 4. Concept of the combined propulsion system

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

combustion engine.

supply with the exhaust gas are possible.

**2.2.1 Parallel power gas turbine supply (variant A)** 

directed to the waste heat boiler in the steam circuit.

parallel, and the steam turbine (variant A)

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 51

exhaust manifold to generate the flow of the scavenge air for supercharging the internal

Present-day designs of turbochargers used in piston engines do not need large amounts of exhaust gas, therefore it seems reasonable to use a power gas turbine complementing the operation of the steam turbine in those cases. Here, two variants of power gas turbine

In this case part of the exhaust gas from the piston engine exhaust manifold supplies the Diesel engine turbocharger. The remaining part of the exhaust gas from the manifold is directed to the gas turbine, bearing the name of the power turbine (PT). The power turbine drives, via the reduction gear, the propeller screw or the electric current generator, thus additionally increasing the power of the entire system. Figure 5 shows a concept of this propulsion system, referred to as parallel power turbine supply. After the expansion in the turbocharger and the power turbine, the exhaust gas flowing from these two turbines is

Fig. 5. Combined system with the Diesel main engine, the power turbine supplied in

In the proposed solution, at low load ranges the amount of the exhaust gas from the main engine is not sufficient to additionally supply the power turbine. In such case a control valve closes the exhaust gas flow to the power turbine, Figure 5. The operation of this valve is controlled by the control system using two signals: the scavenge air pressure signal, and the signal of the propeller shaft angular speed or torque. The waste heat boiler produces the steam which is then used both in the steam turbine and, in case of marine application, to

In the marine low-speed Diesel engines, another portion of energy that can be used along with the exhaust gas energy is a huge amount of so-called waste heat of relatively low temperature. In the low-speed engines the waste heat comprises the following components (with their proportions to the heat delivered to the engine in fuel):


This shows that the amount of the waste heat that remains for our disposal is equal to about 25-30% of the heat delivered in fuel. Part of this heat can be used in the combined circuit with the Diesel engine.

#### **2.1 Energy evaluation of the combined propulsion system**

The adopted concept of the combined ship propulsion system requires energy evaluation, Fig. 4. Formulas defining the system efficiency are derived on the basis of the adopted scheme.

The power of the combined propulsion system is determined by summing up individual powers of system components (the main engine, the power gas turbine, and the steam turbine):

$$\mathbf{N}\_{\rm conbi} = \mathbf{N}\_{\rm D} + \mathbf{N}\_{\rm PT} + \mathbf{N}\_{\rm ST} \tag{1}$$

hence the efficiency of the combined system is:

$$\eta\_{combi} = \frac{\mathcal{N}\_{combi}}{m\_{\rho \text{D}} \cdot \mathcal{V} \mathcal{U} \mu} = \eta\_{\text{D}} \cdot \left( \mathbf{1} + \frac{\mathcal{N}\_{p\_T}}{\mathcal{N}\_{\text{D}}} + \frac{\mathcal{N}\_{\text{ST}}}{\mathcal{N}\_{\text{D}}} \right) \tag{2}$$

and the specific fuel consumption is:

$$b\_{combi} = b\_{e0} \cdot \frac{1}{\left(1 + \frac{N\_{pr}}{N\_D} + \frac{N\_{ST}}{N\_D}\right)} \text{ [g / kWh]} \tag{3}$$

where D, beD- is the efficiency and specific fuel consumption of the main engine.

Relations (2) and (3) show that each additional power in the propulsion system increases the system efficiency and, consequently, decreases the fuel consumption. And the higher the additional power achieved from the utilisation of the heat in the exhaust gas leaving the main engine, the lower the specific fuel consumption. Therefore the maximal available power levels are to be achieved from both the power gas turbine and the steam turbine. The power of the steam turbine mainly depends on the live steam and condenser parameters.

#### **2.2 Variants of the combined ship propulsion systems or marine power plants**

For large powers of low-speed engines, the exhaust gas leaving the engine contains huge amount of heat available for further utilisation. Marine Diesel engines are always supercharged. Portions of the exhaust gas leaving individual cylinders are collected in the exhaust gas collector, where the exhaust gas pressure pexh\_D >pbar is equalised. In standard solutions the constant-pressure turbocharger is supplied with the exhaust gas from the exhaust manifold to generate the flow of the scavenge air for supercharging the internal combustion engine.

Present-day designs of turbochargers used in piston engines do not need large amounts of exhaust gas, therefore it seems reasonable to use a power gas turbine complementing the operation of the steam turbine in those cases. Here, two variants of power gas turbine supply with the exhaust gas are possible.

## **2.2.1 Parallel power gas turbine supply (variant A)**

50 Advances in Gas Turbine Technology

In the marine low-speed Diesel engines, another portion of energy that can be used along with the exhaust gas energy is a huge amount of so-called waste heat of relatively low temperature. In the low-speed engines the waste heat comprises the following components


This shows that the amount of the waste heat that remains for our disposal is equal to about 25-30% of the heat delivered in fuel. Part of this heat can be used in the combined circuit

The adopted concept of the combined ship propulsion system requires energy evaluation, Fig. 4. Formulas defining the system efficiency are derived on the basis of the adopted

The power of the combined propulsion system is determined by summing up individual powers of system components (the main engine, the power gas turbine, and the steam

1 *combi PT ST*

<sup>1</sup> [/ ]

*fD D D N NN m Wu N N*

 

(1 )

Relations (2) and (3) show that each additional power in the propulsion system increases the system efficiency and, consequently, decreases the fuel consumption. And the higher the additional power achieved from the utilisation of the heat in the exhaust gas leaving the main engine, the lower the specific fuel consumption. Therefore the maximal available power levels are to be achieved from both the power gas turbine and the steam turbine. The power of the steam turbine mainly depends on the live steam and condenser parameters.

For large powers of low-speed engines, the exhaust gas leaving the engine contains huge amount of heat available for further utilisation. Marine Diesel engines are always supercharged. Portions of the exhaust gas leaving individual cylinders are collected in the exhaust gas collector, where the exhaust gas pressure pexh\_D >pbar is equalised. In standard solutions the constant-pressure turbocharger is supplied with the exhaust gas from the

*b b g kWh N N N N*

where D, beD- is the efficiency and specific fuel consumption of the main engine.

**2.2 Variants of the combined ship propulsion systems or marine power plants** 

*PT ST D D*

*combi D*

*N NN N combi D PT ST* (1)

(2)

(3)

(with their proportions to the heat delivered to the engine in fuel):

**2.1 Energy evaluation of the combined propulsion system** 

hence the efficiency of the combined system is:

and the specific fuel consumption is:

*ecombi eD*

with the Diesel engine.

scheme.

turbine):

In this case part of the exhaust gas from the piston engine exhaust manifold supplies the Diesel engine turbocharger. The remaining part of the exhaust gas from the manifold is directed to the gas turbine, bearing the name of the power turbine (PT). The power turbine drives, via the reduction gear, the propeller screw or the electric current generator, thus additionally increasing the power of the entire system. Figure 5 shows a concept of this propulsion system, referred to as parallel power turbine supply. After the expansion in the turbocharger and the power turbine, the exhaust gas flowing from these two turbines is directed to the waste heat boiler in the steam circuit.

Fig. 5. Combined system with the Diesel main engine, the power turbine supplied in parallel, and the steam turbine (variant A)

In the proposed solution, at low load ranges the amount of the exhaust gas from the main engine is not sufficient to additionally supply the power turbine. In such case a control valve closes the exhaust gas flow to the power turbine, Figure 5. The operation of this valve is controlled by the control system using two signals: the scavenge air pressure signal, and the signal of the propeller shaft angular speed or torque. The waste heat boiler produces the steam which is then used both in the steam turbine and, in case of marine application, to

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

**3. Power turbine in the combined system** 

\_

*exh D*

corresponding speed indicator.

0,72

1995)

0,74

0,76

0,78

Turbine efficiency

0,8

0,82

0,84

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 53

Calculating the power turbine in the combined system depends on the selected variant of power turbine supply. Usually, piston engine producers do not deliver the exhaust gas temperature in the exhaust manifold (which is equal to the exhaust gas temperature at turbocharger turbine inlet). Instead, they give the exhaust gas temperature at turbocharger turbine outlet (texh\_D). The temperatures of the exhaust gas in the Diesel engine exhaust gas collector are calculated from the turbine power balance, according to the following formula:

\_ o

273,15 -273,15 [ ]

2

(5)

(4)

1

*T g g*

This formula needs the data on turbocharger turbine efficiency changes for partial loads. These data can be obtained from the producer of the turbocharger (as they are rarely made public), Fig. 7, or calculated based on the relation used in steam turbine stage calculations:

2 *<sup>T</sup>*

where - related turbine speed indicator, To- maximal turbine efficiency and the

 

1,2 1,4 1,6 1,8 2 2,2 2,4 2,6 2,8 3 3,2 3,4 3,6 Turbine pressure ratio

Fig. 7. Turbocharger turbine efficiency as a function of scavenge air pressure, acc. to (Schrott,

\_\_\_\_\_\_\_ turbine of S – wheel type \_\_\_\_\_\_\_\_ turbine of R – wheel type

*To* 

*<sup>t</sup> t C*

<sup>1</sup> 1 1

*exh TC*

*T*

*T*

cover the all-ship needs. This system allows for independent operation of the Diesel engine, with the steam turbine or the power turbine switched off. The control system makes it possible to switch off the power turbine thus increasing the power of the turbocharger at partial load, and, on the other hand, direct part of the Diesel engine exhaust gas to supply the power turbine at large load.

Power turbine calculations are based on the Diesel engine parameters, i.e. the temperature of the exhaust gas in the exhaust gas collector, which in turn depends on the engine load and air parameters at the engine inlet. Marine engine producers most often deliver the data on two reference points for the atmospheric air (the ambient reference conditions):


#### **2.2.2 Series power gas turbine supply (variant B)**

In this variant the exhaust gas from the exhaust manifold supplies first the piston engine turbocharger and then the power turbine, Fig.6.

After leaving the exhaust manifold, the exhaust gas expands in the turbocharger to the higher pressure than the atmospheric pressure, which leaves part of the exhaust gas enthalpy drop for utilisation in the power turbine. The exhaust gas leaving the power turbine passes its heat to the steam in the waste heat boiler, thus producing additional power in the steam turbine circuit.

Also in this combined system, the installed control valve makes it possible to switch off the power turbine at partial piston engine loads, thus increasing the power of the turbocharger by expanding the exhaust gas to lower pressure, Fig. 6. Unlike the parallel supply variant, here the entire mass of the exhaust gas from the piston engine manifold flows through the turbocharger. The exhaust gas pressure at the turbocharger outlet is higher than in variant A.

Fig. 6. Combined system with the Diesel main engine, the power turbine supplied in series, and the steam turbine (variant B)

#### **3. Power turbine in the combined system**

52 Advances in Gas Turbine Technology

cover the all-ship needs. This system allows for independent operation of the Diesel engine, with the steam turbine or the power turbine switched off. The control system makes it possible to switch off the power turbine thus increasing the power of the turbocharger at partial load, and, on the other hand, direct part of the Diesel engine exhaust gas to supply

Power turbine calculations are based on the Diesel engine parameters, i.e. the temperature of the exhaust gas in the exhaust gas collector, which in turn depends on the engine load and air parameters at the engine inlet. Marine engine producers most often deliver the data

In this variant the exhaust gas from the exhaust manifold supplies first the piston engine

After leaving the exhaust manifold, the exhaust gas expands in the turbocharger to the higher pressure than the atmospheric pressure, which leaves part of the exhaust gas enthalpy drop for utilisation in the power turbine. The exhaust gas leaving the power turbine passes its heat to the steam in the waste heat boiler, thus producing additional

Also in this combined system, the installed control valve makes it possible to switch off the power turbine at partial piston engine loads, thus increasing the power of the turbocharger by expanding the exhaust gas to lower pressure, Fig. 6. Unlike the parallel supply variant, here the entire mass of the exhaust gas from the piston engine manifold flows through the turbocharger. The exhaust gas pressure at the turbocharger outlet is higher than in variant A.

Fig. 6. Combined system with the Diesel main engine, the power turbine supplied in series,

on two reference points for the atmospheric air (the ambient reference conditions):

ISO Conditions Tropical Conditions

Ambient air temperature [0C] 25 45 Barometric pressure [bar] 1 1

**2.2.2 Series power gas turbine supply (variant B)** 

turbocharger and then the power turbine, Fig.6.

power in the steam turbine circuit.

and the steam turbine (variant B)

the power turbine at large load.

Calculating the power turbine in the combined system depends on the selected variant of power turbine supply. Usually, piston engine producers do not deliver the exhaust gas temperature in the exhaust manifold (which is equal to the exhaust gas temperature at turbocharger turbine inlet). Instead, they give the exhaust gas temperature at turbocharger turbine outlet (texh\_D). The temperatures of the exhaust gas in the Diesel engine exhaust gas collector are calculated from the turbine power balance, according to the following formula:

$$t\_{\rm cub\\_D} = \frac{t\_{\rm cub\\_TC} + 273\,15}{1 - \eta\_r \cdot \left(1 - \frac{1}{\frac{\kappa\_{\rm g} - 1}{\kappa\_{\rm g}}}\right)} \cdot 273\,15 \, [^{\circ}\text{C}] \tag{4}$$

This formula needs the data on turbocharger turbine efficiency changes for partial loads. These data can be obtained from the producer of the turbocharger (as they are rarely made public), Fig. 7, or calculated based on the relation used in steam turbine stage calculations:

$$
\overline{\eta\_{\overline{\nu}}} \equiv \frac{\eta\_{\overline{\nu}}}{\eta\_{\overline{\nu}}} = 2 \cdot \overline{\nu} - \overline{\nu}^2 \tag{5}
$$

where - related turbine speed indicator, To- maximal turbine efficiency and the corresponding speed indicator.

Fig. 7. Turbocharger turbine efficiency as a function of scavenge air pressure, acc. to (Schrott, 1995)

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

**0,550**

**250**

**300**

**350**

**400**

**Temperature [oC]**

**450**

**500**

▲calculated efficiency

function of scavenge air pressure

related to the scavenge air mass flow rate

of Diesel engine load

**0,600**

**0,650**

**0,700**

**TD**

**0,750**

**0,800**

**0,850**

\_\_

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 55

**1,25 1,5 1,75 2 2,25 2,5 2,75 3**

\_\_\_\_\_\_ turbocharger efficiency, acc. to producer \_\_\_\_\_ gas turbine efficiency, acc. to producer ●,

**60 70 80 90 100 110**

\_\_\_\_\_\_ temperature in the Diesel engine exhaust gas collector-calculated curves \_\_\_\_\_\_ exhaust gas temperature at turbocharger outlet – producer's data \_\_\_\_\_\_ Diesel engine exhaust gas mass flow rate

Fig. 10. Sample temperature characteristics of the turbocharger during gas expansion in the turbine to the atmospheric pressure and the related exhaust gas mass flow rates as functions

Fig. 9. Efficiency characteristics of the turbocharger and the turbocharger gas turbine as a

**pD [bar]**

**ND/NDo [%]**

**0,75**

**0,8**

**0,85**

**Relative mass flow**

**0,9**

The turbine speed indicator is defined as:

$$\nu = \frac{\mu}{c\_\ast} = \sqrt{\frac{\mu^2}{2 \cdot H\_\tau}}\tag{6}$$

where u- circumferential velocity on the turbine stage pitch diameter, HT- enthalpy drop in the turbine.

The calculations make use of static characteristics of the turbocharger compressor, with the marked line of cooperation with the Diesel engine, Fig.8.

Figure 9 shows the turbocharger efficiency curves calculated from the relation:

$$
\eta\_{\rm TC} = \eta\_{\rm r} \cdot \eta\_{\rm c} \cdot \eta\_{\rm w} \tag{7}
$$

where T - the turbocharger turbine efficiency is calculated from relation (5), while the compressor efficiency C is calculated from the line of Diesel engine/compressor cooperation, m – mechanical efficiency of the turbocharger, Fig. 8. In the same figure a comparison is made between the calculated turbocharger turbine efficiency with the producer's data as a function of the Diesel engine scavenge pressure. The differences between these curves do not exceed 1,5%.

For the presently available turbocharger efficiency ranges, the amount of the exhaust gas needed for driving the turbocharger turbine is smaller than the entire mass flow rate of the exhaust gas leaving the Diesel engine. Fig. 10 shows sample curves of exhaust gas

Fig. 8. Diesel engine cooperation line against turbocharger compressor characteristics

2

(6)

(7)

2 *s T u u c H*

where u- circumferential velocity on the turbine stage pitch diameter, HT- enthalpy drop in

The calculations make use of static characteristics of the turbocharger compressor, with the

 *TC T C m* 

where T - the turbocharger turbine efficiency is calculated from relation (5), while the compressor efficiency C is calculated from the line of Diesel engine/compressor cooperation, m – mechanical efficiency of the turbocharger, Fig. 8. In the same figure a comparison is made between the calculated turbocharger turbine efficiency with the producer's data as a function of the Diesel engine scavenge pressure. The differences

For the presently available turbocharger efficiency ranges, the amount of the exhaust gas needed for driving the turbocharger turbine is smaller than the entire mass flow rate of the exhaust gas leaving the Diesel engine. Fig. 10 shows sample curves of exhaust gas

Fig. 8. Diesel engine cooperation line against turbocharger compressor characteristics

Figure 9 shows the turbocharger efficiency curves calculated from the relation:

The turbine speed indicator is defined as:

between these curves do not exceed 1,5%.

marked line of cooperation with the Diesel engine, Fig.8.

the turbine.

Fig. 9. Efficiency characteristics of the turbocharger and the turbocharger gas turbine as a function of scavenge air pressure

▲calculated efficiency

\_\_\_\_\_\_ temperature in the Diesel engine exhaust gas collector-calculated curves \_\_\_\_\_\_ exhaust gas temperature at turbocharger outlet – producer's data \_\_\_\_\_\_ Diesel engine exhaust gas mass flow rate related to the scavenge air mass flow rate

Fig. 10. Sample temperature characteristics of the turbocharger during gas expansion in the turbine to the atmospheric pressure and the related exhaust gas mass flow rates as functions of Diesel engine load

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

**290**

**310**

**330**

**350**

**370**

**390**

**Temperature [ oC ]**

**410**

**430**

**450**

**470**

**490**

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 57

**60 70 80 90 100 110**

\_\_\_\_\_expansion ratio in the turbocharger turbine (standard arrangement - without power turbine) \_ \_ \_ expansion ratio in the turbocharger turbine with power turbine \_\_\_\_\_\_exhaust gas temperature in the Diesel engine exhaust gas collector \_\_\_\_exhaust gas temperature at turbocharger outlet without power

Fig. 11. Changes of temperature and expansion ratio of the turbocharger in the combined

The mass flow rate of the exhaust gas needed by the turbocharger is calculated from the

1

*g g*

1 *TC T exh D g*

*a a p*

1

*m T c*

The exhaust gas expanding in the power turbine has the inlet and outlet pressures identical to those of the exhaust gas flowing through the turbocharger. The power of the power

> *N mH PT m PT PT PT*

where m- mechanical efficiency of the power turbine, HPT – iso-entropic enthalpy drop in

The power turbine efficiency PT is assumed in the same way as for the turbocharger turbine, Fig. 9, or using the relation (5). In the shipbuilding, the gas turbines used in combined Diesel engine systems with power turbines are those adopted from turbochargers.

 

*a a m T c*

<sup>1</sup> <sup>1</sup>

*C*

\_

*TC*

turbine \_ \_ \_ exhaust gas temperature at turbocharger outlet with power turbine

system with series power turbine supply (variant B)

turbine is given by the relation:

the power turbine.

turbocharger power balance using the following formula:

*m*

**ND/NDo [%]**

**1,2**

(10.1)

(11)

**1,6**

**2**

**Expansion ratio**

**2,4**

**2,8**

temperature changes in the engine manifold (calculated using the relation (4)) and the exhaust gas temperature at the turbocharger outlet (according to the data delivered by the producer) as functions of engine load, when the standard internal combustion engine exhaust gas is expanded to the barometric pressure. The figure also shows the Diesel engine exhaust gas flow rate related to the scavenge air flow rate, as a function of the engine load. This high efficiency of the turbocharger provides opportunities for installing a power gas turbine connected in parallel with the turbocharger (variant A).

The turbocharger power balance indicates that in the power gas turbine we can utilise between 10 and 24% of the flow rate of the exhaust gas leaving the exhaust manifold of the piston engine. The power gas turbine can be switched on when the main engine power output exceeds 60%. For lower power outputs the entire exhaust gas flow leaving the Diesel engine is to be used for driving the turbocharger.

In variant B of the combined system with the power turbine, the turbocharger is connected in series with the power gas turbine. Here, the entire amount of the exhaust gas flows through the turbocharger turbine. Due to the excess of the power needed for driving the turbocharger, the final expansion pressure at turbocharger turbine output can be higher than the exhaust gas pressure at waste heat boiler inlet. In this case the expansion ratio in the turbocharger turbine is given by the relation:

$$\pi\_r = \left[ \frac{1}{1 - \frac{1}{\eta\_{r\text{c}}} \cdot \frac{m\_s}{m\_\odot} \cdot \frac{c\_s}{c\_s} \cdot \frac{t\_s}{t\_{cch\_\text{-}D}} \cdot \left(\pi\_c \overset{\text{NaCl}}{^{\text{NaCl}}}\right)}\right]^{\frac{\kappa\_\text{g}}{\kappa\_\text{g}-1}}\tag{8}$$

where: C- compression ratio of the turbocharger compressor. The exhaust gas temperature at turbocharger outlet is calculated from the formula:

$$\left| t\_{\rm{ch}\_{\rm{-}}\rm{TC}} = \left( t\_{\rm{ch}\_{\rm{-}}D} + 273, 15 \right) \cdot \left| 1 - \eta\_{\rm{I}} \left( 1 - \frac{1}{\frac{\frac{\kappa\_{\rm{S}}}{\kappa\_{\rm{S}}}}{\pi\_{\rm{S}}}} \right) \right| - 273, 15 \text{ [}^{\circ}\text{C]} \tag{9}$$

Figure 11 shows sample curves of temperature, compression and expansion rate changes in the turbocharger for variant B: series power turbine supply.

This case provides opportunities for utilising the enthalpy drop of the expanding exhaust gas in the power turbine. The operation of the power turbine is possible when the Diesel engine power exceeds 60%.

#### **3.1 Power turbine in parallel supply system (variant A)**

The power turbine (Fig.5) is supplied with the exhaust gas from the exhaust manifold. The exhaust gas mass flow rate mPT and temperature texh\_D are identical as those at turbocharger outlet: the mass flow rate of the exhaust gas flowing through the power turbine results from the difference between the mass flow rate of the Diesel engine exhaust gas and of that expanding in the turbocharger:

$$m\_{\rm TD} = m\_{\rm a} \cdot (1 - \overline{m}) + m\_{\rm f^{\odot}} \tag{10}$$

temperature changes in the engine manifold (calculated using the relation (4)) and the exhaust gas temperature at the turbocharger outlet (according to the data delivered by the producer) as functions of engine load, when the standard internal combustion engine exhaust gas is expanded to the barometric pressure. The figure also shows the Diesel engine exhaust gas flow rate related to the scavenge air flow rate, as a function of the engine load. This high efficiency of the turbocharger provides opportunities for installing a power gas

The turbocharger power balance indicates that in the power gas turbine we can utilise between 10 and 24% of the flow rate of the exhaust gas leaving the exhaust manifold of the piston engine. The power gas turbine can be switched on when the main engine power output exceeds 60%. For lower power outputs the entire exhaust gas flow leaving the Diesel

In variant B of the combined system with the power turbine, the turbocharger is connected in series with the power gas turbine. Here, the entire amount of the exhaust gas flows through the turbocharger turbine. Due to the excess of the power needed for driving the turbocharger, the final expansion pressure at turbocharger turbine output can be higher than the exhaust gas pressure at waste heat boiler inlet. In this case the expansion ratio in

1

 

*aa a*

 

*TC D g exh D*

The exhaust gas temperature at turbocharger outlet is calculated from the formula:

\_ \_ 1

*mct*

1

(8)

(9)

*g g*

(1 ) *m m mm TD a fD* (10)

 

1

 

*C*

*T g g*

 

 

\_

*<sup>a</sup> mc t <sup>a</sup>*

<sup>o</sup>

<sup>1</sup> 273,15 1 1 273,15 [ C] *exh TC exh D <sup>T</sup>*

Figure 11 shows sample curves of temperature, compression and expansion rate changes in

This case provides opportunities for utilising the enthalpy drop of the expanding exhaust gas in the power turbine. The operation of the power turbine is possible when the Diesel

The power turbine (Fig.5) is supplied with the exhaust gas from the exhaust manifold. The exhaust gas mass flow rate mPT and temperature texh\_D are identical as those at turbocharger outlet: the mass flow rate of the exhaust gas flowing through the power turbine results from the difference between the mass flow rate of the Diesel engine exhaust gas and of that

turbine connected in parallel with the turbocharger (variant A).

<sup>1</sup> <sup>1</sup>

engine is to be used for driving the turbocharger.

the turbocharger turbine is given by the relation:

*t t*

engine power exceeds 60%.

expanding in the turbocharger:

*T*

where: C- compression ratio of the turbocharger compressor.

the turbocharger for variant B: series power turbine supply.

**3.1 Power turbine in parallel supply system (variant A)** 

\_\_\_\_\_expansion ratio in the turbocharger turbine (standard arrangement - without power turbine) \_ \_ \_ expansion ratio in the turbocharger turbine with power turbine \_\_\_\_\_\_exhaust gas temperature in the Diesel engine exhaust gas collector \_\_\_\_exhaust gas temperature at turbocharger outlet without power turbine \_ \_ \_ exhaust gas temperature at turbocharger outlet with power turbine

Fig. 11. Changes of temperature and expansion ratio of the turbocharger in the combined system with series power turbine supply (variant B)

The mass flow rate of the exhaust gas needed by the turbocharger is calculated from the turbocharger power balance using the following formula:

$$\mathbf{1} - \frac{1}{m\_{\text{rc}}} \underbrace{\frac{1}{\frac{\kappa\_{\text{g}} - 1}{\kappa\_{\text{g}}}}}\_{\begin{subarray}{c} \frac{\kappa\_{\text{g}} - 1}{\kappa\_{\text{g}}} \\ \pi\_{\text{c}} \end{subarray}} \underbrace{\frac{1}{\kappa\_{\text{g}} - 1}}\_{\begin{subarray}{c} \frac{\kappa\_{\text{g}} - 1}{\kappa\_{\text{g}}} \\ -1 \end{subarray}} \cdot \frac{T\_{\text{c}ch\_{-}D}}\_{\begin{subarray}{c} T\_{\text{s}} \end{subarray}} \cdot \frac{c\_{\text{g}}}{c\_{\text{r}}} \cdot \eta\_{\text{rc}} \tag{10.1}$$

The exhaust gas expanding in the power turbine has the inlet and outlet pressures identical to those of the exhaust gas flowing through the turbocharger. The power of the power turbine is given by the relation:

$$\mathcal{N}\_{\rm PT} = \boldsymbol{\eta}\_{\rm m} \cdot \boldsymbol{\eta}\_{\rm PT} \cdot \boldsymbol{m}\_{\rm pr} \cdot \boldsymbol{H}\_{\rm pr} \tag{11}$$

where m- mechanical efficiency of the power turbine, HPT – iso-entropic enthalpy drop in the power turbine.

The power turbine efficiency PT is assumed in the same way as for the turbocharger turbine, Fig. 9, or using the relation (5). In the shipbuilding, the gas turbines used in combined Diesel engine systems with power turbines are those adopted from turbochargers.

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

that of the exhaust gas leaving the turbocharger, Fig. 13.

*t t*

the same way as in variant A.

engine load.

**300**

**320**

**340**

**temperature [ oC ]**

**360**

**380**

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 59

where tinl\_PT - exhaust gas temperature at the power turbine inlet, PT– expansion ratio in the power turbine , PT- power turbine efficiency. The power turbine efficiency is assumed in

In formula (12) the exhaust gas temperature at the power turbine inlet is assumed equal to

Figure 13 also shows the expansion ratio, the power of the power turbine, and the exhaust gas temperatures at the turbocharger and the power turbine outlets for partial engine loads. The power turbine in this variant increases the power of the combined system by 3% to 9% with respect to that of a standard engine. The turbine power increases with increasing Diesel

**60 70 80 90 100 110**

Fig. 13. Parameters of series supplied power turbine as functions of the main engine load -

The analysis of the two examined variants shows that the power of the combined system increases depending on the Diesel engine load. For both variants the power turbine can be

**\_\_\_\_** temperature at turbocharger outlet **\_\_\_\_** temperature at power turbine outlet **\_\_\_\_** expansion ratio in power turbine **\_\_\_\_** related power of power turbine

variant B (calculations for tropical conditions)

**3.3 Comparing the two power turbine supply variants** 

<sup>1</sup> 273,15 1 1 273,15[ ] *<sup>o</sup>*

*PT g g*

 

*C*

(13)

**ND/NDo [ % ]**

**0**

**0,2**

**0,4**

**0,6**

**0,8**

**Expansion ratio [-], Relative power x10 [%]**

**1**

**1,2**

**1,4**

 

The exhaust gas temperature at the power turbine output is calculated from the formula:

\_ \_ <sup>1</sup>

*exh PT inl PT PT*

The power turbine system calculations show that the exhaust gas temperature at the power turbine outlet is slightly higher than that at the turbocharger outlet, Fig.12. The increase of the main engine load results in the increase of both the exhaust gas temperature in the exhaust gas collector and the mass flow rate of the exhaust gas flowing through the power turbine. The increase in power of the combined system with additional power turbine ranges from about 2% for Diesel engine loads of an order of 70% up to over 8% for maximal loads, Fig.12.

exhaust gas mass flow rate in power turbine \_\_\_\_\_ related power turbine power

Fig. 12. Parameters of parallel supplied power turbine as functions of the main engine load – variant A (calculations for tropical conditions)

When the Diesel engine power is lower than 60-70% of the nominal value the entire exhaust gas flow from the exhaust manifold is directed to the turbocharger drive. In this case the control system closes the valve controlling the exhaust gas flow to the power turbine, Fig. 5.

#### **3.2 Power turbine in series supply system (variant B)**

In this variant the power turbine is supplied with the full amount of the exhaust gas leaving the Diesel engine exhaust manifold. The power turbine is installed after the turbocharger. The exhaust gas pressure at the power turbine inlet depends on the pressure of the exhaust gas leaving the turbocharger turbine, Fig.11.

In this case the power of the power turbine is calculated as:

$$\mathcal{N}\_{PT} = \eta\_{PT} \cdot m\_D \cdot c\_g \cdot t\_{\text{int\\_}^{\text{-}T}} \cdot \left(1 - \frac{1}{\frac{\frac{\mathcal{N}\_{\text{-}S}^{-1}}{\mathcal{N}\_{\text{S}}}}}\right) \tag{12}$$

The power turbine system calculations show that the exhaust gas temperature at the power turbine outlet is slightly higher than that at the turbocharger outlet, Fig.12. The increase of the main engine load results in the increase of both the exhaust gas temperature in the exhaust gas collector and the mass flow rate of the exhaust gas flowing through the power turbine. The increase in power of the combined system with additional power turbine ranges from about 2% for Diesel engine loads of an order of 70% up to over 8% for maximal

**60 70 80 90 100 110**

Fig. 12. Parameters of parallel supplied power turbine as functions of the main engine load –

When the Diesel engine power is lower than 60-70% of the nominal value the entire exhaust gas flow from the exhaust manifold is directed to the turbocharger drive. In this case the control system closes the valve controlling the exhaust gas flow to the power

In this variant the power turbine is supplied with the full amount of the exhaust gas leaving the Diesel engine exhaust manifold. The power turbine is installed after the turbocharger. The exhaust gas pressure at the power turbine inlet depends on the pressure of the exhaust

<sup>1</sup> <sup>1</sup> *PT PT D g inl PT*

*N mct* 

\_ 1

*PT g g*

 

 

\_\_\_\_\_ temperature at turbocharger outlet \_\_\_\_\_ temperature at power turbineoutlet \_\_\_\_\_ related

exhaust gas mass flow rate in power turbine \_\_\_\_\_ related power turbine power

variant A (calculations for tropical conditions)

gas leaving the turbocharger turbine, Fig.11.

**3.2 Power turbine in series supply system (variant B)** 

In this case the power of the power turbine is calculated as:

**ND/NDo [%]**

(12)

**0**

**5**

**10**

**Relative gas flow, Relative power [%]**

**15**

**20**

loads, Fig.12.

**290**

turbine, Fig. 5.

**300**

**310**

**320**

**Ttemperature [oC]**

**330**

**340**

where tinl\_PT - exhaust gas temperature at the power turbine inlet, PT– expansion ratio in the power turbine , PT- power turbine efficiency. The power turbine efficiency is assumed in the same way as in variant A.

In formula (12) the exhaust gas temperature at the power turbine inlet is assumed equal to that of the exhaust gas leaving the turbocharger, Fig. 13.

The exhaust gas temperature at the power turbine output is calculated from the formula:

$$\mathbf{t}\_{\text{e}\_{\text{e}},\text{}\_{\text{}},\text{}^{\text{T}}} = \left(\mathbf{t}\_{\text{in}\_{\text{}},\text{}^{\text{T}}} + 273, 15\right) \cdot \left[\mathbf{1} - \eta\_{\text{T}} \begin{pmatrix} \mathbf{1} \\ \mathbf{1} - \frac{\mathbf{1}}{\frac{\mathbf{x}\_{\text{S}} - \mathbf{1}}{\mathbf{x}\_{\text{S}}}} \\ \pi\_{\text{T}} \end{pmatrix} \right] - 273, 15[^{\circ}\text{C}] \tag{13}$$

Figure 13 also shows the expansion ratio, the power of the power turbine, and the exhaust gas temperatures at the turbocharger and the power turbine outlets for partial engine loads. The power turbine in this variant increases the power of the combined system by 3% to 9% with respect to that of a standard engine. The turbine power increases with increasing Diesel engine load.

**\_\_\_\_** temperature at turbocharger outlet **\_\_\_\_** temperature at power turbine outlet **\_\_\_\_** expansion ratio in power turbine **\_\_\_\_** related power of power turbine

Fig. 13. Parameters of series supplied power turbine as functions of the main engine load variant B (calculations for tropical conditions)

#### **3.3 Comparing the two power turbine supply variants**

The analysis of the two examined variants shows that the power of the combined system increases depending on the Diesel engine load. For both variants the power turbine can be

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

**0**

**1**

**2**

**3**

**4**

**5**

**Relative power [%]**

**6**

**7**

**8**

**9**

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 61

**60 70 80 90 100 110**

**\_\_\_\_** parallel power turbine supply (variant A) **\_\_\_\_** series power turbine supply (variant B)

1-Waste Heat Boiler 2-Superheater 3- Evaporator 4-Ekonomizer 5-Boiler drum 6-Steam turbine 7-

Condenser 8-Deaerator 9-Feed water pump 10-Condensate pump

Fig. 15. Flow Diagram of the Single Pressure System

Fig. 14. Powers of the power turbine as functions of main engine load

**ND/NDo [%]**

**10**

**15**

**20**

 **Difference power [%]**

**25**

**30**

used after exceeding about 65% of the Diesel engine power. The exhaust gas leaving the power turbine is directed to the waste heat boiler, where together with steam turbine it can additionally increase the overall power of the combined system.

In both cases the temperatures of the exhaust gas leaving the power turbine are comparable. The exhaust gas pressure at power turbine outlet depends on the losses generated when the gas flows through the waste heat boiler and outlet silencers. Following practical experience, the exhaust gas back pressure is assumed higher than the barometric pressure by 300 mmWC, i.e. about 3%. Taking into account powers of the power turbines for the above variants, Fig. 14, it shows that for the same Diesel engine parameters the series supply of the power turbine results in higher turbine power. For lower loads, the power of the series supplied power turbine increases, compared to the parallel supply variant.

## **4. Steam turbine circuit**

The combined system makes use of the waste heat from the Diesel engine. In modern Diesel engines the temperatures of the waste heat are at the advantageous levels for the steam turbine circuit. This circuit makes use of water that can be utilised in a low-temperature process. Adding the steam circuit to the combined Diesel engine/power gas turbine system provides good opportunities for increasing the power of the combined system, and consequently, also the system efficiency, see formula (2).

In the examined combined system the exhaust gas leaving the turbocharger and the power turbine (variant A, Fig. 5) or only the power turbine (variant B, Fig. 6) flows to the waste heat boiler where it is used for producing superheated steam for driving the steam turbine.

The mass flow rate of the exhaust gas reaching the waste heat boiler is equal to that leaving the Diesel engine exhaust gas collector. The exhaust gas temperature at waste heat boiler inlet depends on the adopted solution of power turbine supply. For variant A with parallel supply it is calculated from the balance of mixing of the gases leaving the turbocharger and the power turbine:

$$\mathbf{t}\_{\rm id\\_8} = \frac{m\_{\rm TC\\_i} \cdot \dot{\mathbf{t}}\_{\rm cch\\_TC} + m\_{\rm PT\\_i} \cdot \dot{\mathbf{t}}\_{\rm cch\\_PT}}{m\_{\rm D} \cdot \mathbf{c}\_{\rm g}} - 273\,\mathrm{J}\,\mathrm{5}\,\,\left[\,^\circ\mathrm{C}\right] \tag{14}$$

while for the series power turbine supply (variant B) it is assumed equal to that at the power turbine outlet, formula (13).

In combined steam turbine systems for small power ranges and low live steam temperatures the single pressure systems are used, Fig. 15, (Kehlhofer, 1991).

Such system consists of a single-pressure waste heat boiler, a condensing steam turbine, a water-cooled condenser, and a single stage feed water preheater in the deaerator.

The main disadvantage of the systems of this type is poor utilisation of the heat contained in the exhaust gas (the waste heat energy). The steam superheater is relatively large, as the entire mass of the steam produced by the boiler flows through it. However, costs of this steam system are the lowest, as poor utilisation of the exhaust gas energy results in high temperature of the exhaust gas leaving the boiler. The deaerator is supplied with the steam extracted from the steam turbine. The application of the single pressure system does not secure optimal utilisation of the exhaust gas energy.

used after exceeding about 65% of the Diesel engine power. The exhaust gas leaving the power turbine is directed to the waste heat boiler, where together with steam turbine it can

In both cases the temperatures of the exhaust gas leaving the power turbine are comparable. The exhaust gas pressure at power turbine outlet depends on the losses generated when the gas flows through the waste heat boiler and outlet silencers. Following practical experience, the exhaust gas back pressure is assumed higher than the barometric pressure by 300 mmWC, i.e. about 3%. Taking into account powers of the power turbines for the above variants, Fig. 14, it shows that for the same Diesel engine parameters the series supply of the power turbine results in higher turbine power. For lower loads, the power of the series

The combined system makes use of the waste heat from the Diesel engine. In modern Diesel engines the temperatures of the waste heat are at the advantageous levels for the steam turbine circuit. This circuit makes use of water that can be utilised in a low-temperature process. Adding the steam circuit to the combined Diesel engine/power gas turbine system provides good opportunities for increasing the power of the combined system, and

In the examined combined system the exhaust gas leaving the turbocharger and the power turbine (variant A, Fig. 5) or only the power turbine (variant B, Fig. 6) flows to the waste heat boiler where it is used for producing superheated steam for driving the steam

The mass flow rate of the exhaust gas reaching the waste heat boiler is equal to that leaving the Diesel engine exhaust gas collector. The exhaust gas temperature at waste heat boiler inlet depends on the adopted solution of power turbine supply. For variant A with parallel supply it is calculated from the balance of mixing of the gases leaving the turbocharger and

> \_ \_ \_ 273,15 [ ] *TC exh TC PT exh PT <sup>o</sup>*

while for the series power turbine supply (variant B) it is assumed equal to that at the power

In combined steam turbine systems for small power ranges and low live steam temperatures

Such system consists of a single-pressure waste heat boiler, a condensing steam turbine, a

The main disadvantage of the systems of this type is poor utilisation of the heat contained in the exhaust gas (the waste heat energy). The steam superheater is relatively large, as the entire mass of the steam produced by the boiler flows through it. However, costs of this steam system are the lowest, as poor utilisation of the exhaust gas energy results in high temperature of the exhaust gas leaving the boiler. The deaerator is supplied with the steam extracted from the steam turbine. The application of the single pressure system does not

(14)

*D g mi mi t C m c*

water-cooled condenser, and a single stage feed water preheater in the deaerator.

additionally increase the overall power of the combined system.

consequently, also the system efficiency, see formula (2).

*inl B*

secure optimal utilisation of the exhaust gas energy.

the single pressure systems are used, Fig. 15, (Kehlhofer, 1991).

**4. Steam turbine circuit** 

turbine.

the power turbine:

turbine outlet, formula (13).

supplied power turbine increases, compared to the parallel supply variant.

Fig. 14. Powers of the power turbine as functions of main engine load

1-Waste Heat Boiler 2-Superheater 3- Evaporator 4-Ekonomizer 5-Boiler drum 6-Steam turbine 7- Condenser 8-Deaerator 9-Feed water pump 10-Condensate pump

Fig. 15. Flow Diagram of the Single Pressure System

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

**4.1 Limits for steam circuit parameters** 

temperature is tFW > 140-145oC (Kehlhofer, 1991).

0

5

10

15

20

**Condenser pressure [kPa]**

25

30

35

40

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 63

The limits for the values of the steam circuit parameters result from strength and technical requirements concerning the durability of particular system components, but also from design and economic restrictions. The difference between the exhaust gas temperature and the live steam temperature, t, for waste heat boilers used in shipbuilding is assumed as t = 10-15oC, according to (MAN, 1985; Kehlhofer, 1991). The "pitch point" value recommended by MAN B&W (MAN, 1985) for marine boilers is t = 8-12oC. The limiting dryness factor x of the steam downstream of the steam turbine is assumed as xlimit=0,86-0,88. For marine condensers cooled with sea water, MAN recommends the condenser pressure pK=0,065 bar. This pressure depends on the B&W (MAN, 1985) temperature of the cooling medium in the condenser. Figure 17 shows the dependence of the condenser pressure on the cooling medium temperature. The temperature of the boiler feed water is of high importance for the life time of the feed water heater in the boiler. The value of this temperature is connected with a so-called exhaust gas dew-point temperature. Below this temperature the water condensates on heater tubes and reacts with the sulphur trioxide SO3 producing the sulphuric acid, which is the source of low-temperature corrosion. That is why boiler producers give minimal feed water temperatures below which boiler operation is highly not recommended. The dew-point temperature is connected with the content of sulphur in the fuel and depends on the excess air coefficient in the piston engine. Figure 18 shows the dew-point temperature as the function of: sulphur content in the fuel, SO2 conversion to SO3, and the excess air coefficient in the engine. In inland power installations burning fuels with sulphur content higher than 2%, the recommended level of feed water


**\_\_\_\_** Fresh Water Cooling **\_\_\_\_** Wet Cooling Tower **\_\_\_\_** Direct Air Condensation

Fig. 17. Condenser pressure as a function of temperature of the cooling medium

**C]**

Those steam turbine systems frequently make use of an additional low-pressure evaporator, Fig. 16, which leads not only to more intensive utilisation of the waste heat contained in the exhaust gas, but also to better thermodynamic use of the low-pressure steam.

In this solution the high pressure superheater is relatively small, compared to the single pressure boiler. The deaerator is heated with the saturated steam from the low-pressure evaporator. The power of the main high-pressure feeding pump is also smaller. The excess steam from the low-pressure evaporator can be used for supplying the low-pressure part of the steam turbine, thus increasing its power, or, alternatively, for covering all-ship needs.

Figure 16 shows possible use of the temperature waste heat from the scavenge air cooler, the lubricating oil cooler, and from the jacket water cooler in the low-pressure water preheater.

The additional low-pressure exchanger in the steam circuit, Fig. 16, makes it possible to increase the temperature of the water in the deaerator. Higher water temperature is required due to the presence of sulphur in the fuel (water dew-point in the exhaust gas) – it is favourable for systems fed with a high sulphur content fuel. If the temperature of the feedwater is low when the system is fed with fuel without sulphur, the heat exchanger 14 in Fig. 16 is not necessary and the waste heat from the coolers can be used in the deaerator. For a low feedwater temperature the deaerator works at the pressure below atmospheric (under the vacuum).

1-Waste Heat Boiler 2-High pressure superheater 3- High pressure evaporator 4- High pressure economizer 5- High pressure boiler drum 6 -Steam turbine 7-Condenser 8- Deaerator 9-High pressure feed water pump 10-Condensate pump 11-Low pressure feed pump 12-Low pressure evaporator 13- Low pressure boiler drum 14-Low pressure pre-heater

Fig. 16. Flow Diagram for a Two – Pressure System

### **4.1 Limits for steam circuit parameters**

62 Advances in Gas Turbine Technology

Those steam turbine systems frequently make use of an additional low-pressure evaporator, Fig. 16, which leads not only to more intensive utilisation of the waste heat contained in the

In this solution the high pressure superheater is relatively small, compared to the single pressure boiler. The deaerator is heated with the saturated steam from the low-pressure evaporator. The power of the main high-pressure feeding pump is also smaller. The excess steam from the low-pressure evaporator can be used for supplying the low-pressure part of the steam turbine, thus increasing its power, or, alternatively, for covering all-ship needs. Figure 16 shows possible use of the temperature waste heat from the scavenge air cooler, the lubricating oil cooler, and from the jacket water cooler in the low-pressure water pre-

The additional low-pressure exchanger in the steam circuit, Fig. 16, makes it possible to increase the temperature of the water in the deaerator. Higher water temperature is required due to the presence of sulphur in the fuel (water dew-point in the exhaust gas) – it is favourable for systems fed with a high sulphur content fuel. If the temperature of the feedwater is low when the system is fed with fuel without sulphur, the heat exchanger 14 in Fig. 16 is not necessary and the waste heat from the coolers can be used in the deaerator. For a low feedwater temperature the deaerator works at the pressure below atmospheric

1-Waste Heat Boiler 2-High pressure superheater 3- High pressure evaporator 4- High pressure economizer 5- High pressure boiler drum 6 -Steam turbine 7-Condenser 8- Deaerator 9-High pressure feed water pump 10-Condensate pump 11-Low pressure feed pump 12-Low pressure evaporator 13-

Low pressure boiler drum 14-Low pressure pre-heater Fig. 16. Flow Diagram for a Two – Pressure System

exhaust gas, but also to better thermodynamic use of the low-pressure steam.

heater.

(under the vacuum).

The limits for the values of the steam circuit parameters result from strength and technical requirements concerning the durability of particular system components, but also from design and economic restrictions. The difference between the exhaust gas temperature and the live steam temperature, t, for waste heat boilers used in shipbuilding is assumed as t = 10-15oC, according to (MAN, 1985; Kehlhofer, 1991). The "pitch point" value recommended by MAN B&W (MAN, 1985) for marine boilers is t = 8-12oC. The limiting dryness factor x of the steam downstream of the steam turbine is assumed as xlimit=0,86-0,88. For marine condensers cooled with sea water, MAN recommends the condenser pressure pK=0,065 bar. This pressure depends on the B&W (MAN, 1985) temperature of the cooling medium in the condenser. Figure 17 shows the dependence of the condenser pressure on the cooling medium temperature. The temperature of the boiler feed water is of high importance for the life time of the feed water heater in the boiler. The value of this temperature is connected with a so-called exhaust gas dew-point temperature. Below this temperature the water condensates on heater tubes and reacts with the sulphur trioxide SO3 producing the sulphuric acid, which is the source of low-temperature corrosion. That is why boiler producers give minimal feed water temperatures below which boiler operation is highly not recommended. The dew-point temperature is connected with the content of sulphur in the fuel and depends on the excess air coefficient in the piston engine. Figure 18 shows the dew-point temperature as the function of: sulphur content in the fuel, SO2 conversion to SO3, and the excess air coefficient in the engine. In inland power installations burning fuels with sulphur content higher than 2%, the recommended level of feed water temperature is tFW > 140-145oC (Kehlhofer, 1991).

**\_\_\_\_** Fresh Water Cooling **\_\_\_\_** Wet Cooling Tower **\_\_\_\_** Direct Air Condensation Fig. 17. Condenser pressure as a function of temperature of the cooling medium

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

system within the entire range of the main engine load.

with series power turbine supply.

steam turbine without this exchanger.

**5.1 Power range of combined systems** 

**5.2 Efficiency of combined systems** 

piston engine load.

**5. Conclusions** 

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 65

the sub-area where the earlier discussed limits imposed on the steam system are met. The use of the steam system with the waste heat boiler increases the power of the propulsion

Adding a steam turbine to the Diesel engine system increases the power of the propulsion system by NST/ND = 6,5 – 7,5% for main engine loads ranging from 90 to 100%. The power of the steam turbine for both examined variants of power turbine supply are comparable, and slightly higher power, by about 2-4%, is obtained by the steam turbine in the variant

The analysis of the system with an additional exchanger utilising the low-temperature waste heat from the Diesel engine to heat the condensate from the condenser before the deaerator, Fig.16, shows that the steam turbine power increases by 7- 11% with respect to that of the

The requirements concerning the waste heat boiler refer to low loss of the exhaust gas flow (which reduces the final expansion pressure in the power turbine) and small temperature concentrations (pitch points) in the boiler evaporators. There is a remarkable impact of the sulphur content in the fuel on the permissible exhaust gas temperature and the lower feed water temperature limit. In the steam turbine circuit, a minimal number of exchangers should be used (optimally: none). The optimal parameters of this circuit also depend on the

It is possible to implement a combined system consisting of a Diesel engine as the leading engine, a power gas turbine, and a steam turbine circuit utilising the heat contained in the Diesel engine exhaust gas. Such systems can reveal thermodynamic efficiencies comparable

Depending on the adopted variant and the main engine load, the use of the combined system makes it possible to increase the power of the power plant by 7 to 15 % with respect to the conventional power plant burning the same rate of fuel. Additional power is obtained by the system due to the recovery of the energy contained in the exhaust gas leaving the piston internal combustion engine. Thus the combined system decreases the specific fuel

In the examined systems the power of the steam turbine is higher than that of the power

The use of the combined system for ship propulsion increases the efficiency of the propulsion system, and decreases the specific fuel consumption. Additionally, it increases

Like the power, the efficiency of the combined system increases with respect to the conventional power plant by 7 to 15% reaching the level of 53 - 56% for maximal power ranges. These efficiency levels are comparable with the combined systems based on the steam/gas turbines, Fig. 1. For partial loads the efficiency curves of the combined system

with combined gas turbine circuits connected with steam turbines.

consumption by 6,4 – 12,8 % compared to the conventional power plant.

the propulsion power without additional fuel consumption.

turbine by 6-29 %, depending on the system variant and the main engine load.

In marine propulsion (MAN, 1985) recommends that the feed water temperature should not be lower than 120oC when the sulphur content is higher than 2%. This is justified by the fact that the outer surface of the heater tubes on the exhaust gas side has the temperature higher by 8-15 oC than the feed water temperature, and that the materials used in those heaters reveal enhanced resistance to acid corrosion.

The exhaust gas temperature at the boiler outlet is assumed higher by 15-20oC than the feed water temperature, i.e. texh > tFW + (15 – 20oC).

Each ship burning heavy fuel in its power plant uses the mass flow rate mSS of the saturated steam taken from the waste heat boiler for fuel pre-heating and all-ship purposes. According to the recommendations (MAN, 1985) the pressure of the steam used for these purposes should range between pSS = 7-9 bar. This pressure is also assumed equal to the pressure in the boiler low-pressure circuit. The back temperature of the above steam flow in the heat box is within 50 – 60oC.

Fig. 18. Acid dew-point as a function of the sulphur content in the fuel and the excess air coefficient

#### **4.2 Optimising the steam circuit**

Optimisation of the steam system is to be done in such a way so as to reach the maximal possible utilisation of the heat contained in the exhaust gas. In this sense the optimisation is reduced to selecting the steam circuit parameters for which the steam turbine reaches the highest power. The area of search for optimal steam circuit parameters is to be narrowed to the sub-area where the earlier discussed limits imposed on the steam system are met. The use of the steam system with the waste heat boiler increases the power of the propulsion system within the entire range of the main engine load.

Adding a steam turbine to the Diesel engine system increases the power of the propulsion system by NST/ND = 6,5 – 7,5% for main engine loads ranging from 90 to 100%. The power of the steam turbine for both examined variants of power turbine supply are comparable, and slightly higher power, by about 2-4%, is obtained by the steam turbine in the variant with series power turbine supply.

The analysis of the system with an additional exchanger utilising the low-temperature waste heat from the Diesel engine to heat the condensate from the condenser before the deaerator, Fig.16, shows that the steam turbine power increases by 7- 11% with respect to that of the steam turbine without this exchanger.

The requirements concerning the waste heat boiler refer to low loss of the exhaust gas flow (which reduces the final expansion pressure in the power turbine) and small temperature concentrations (pitch points) in the boiler evaporators. There is a remarkable impact of the sulphur content in the fuel on the permissible exhaust gas temperature and the lower feed water temperature limit. In the steam turbine circuit, a minimal number of exchangers should be used (optimally: none). The optimal parameters of this circuit also depend on the piston engine load.

## **5. Conclusions**

64 Advances in Gas Turbine Technology

In marine propulsion (MAN, 1985) recommends that the feed water temperature should not be lower than 120oC when the sulphur content is higher than 2%. This is justified by the fact that the outer surface of the heater tubes on the exhaust gas side has the temperature higher by 8-15 oC than the feed water temperature, and that the materials used in those heaters

The exhaust gas temperature at the boiler outlet is assumed higher by 15-20oC than the feed

Each ship burning heavy fuel in its power plant uses the mass flow rate mSS of the saturated steam taken from the waste heat boiler for fuel pre-heating and all-ship purposes. According to the recommendations (MAN, 1985) the pressure of the steam used for these purposes should range between pSS = 7-9 bar. This pressure is also assumed equal to the pressure in the boiler low-pressure circuit. The back temperature of the above steam flow in the heat

0,01 0,1 1 10 **Sulphur content in the fuel [%]**

Fig. 18. Acid dew-point as a function of the sulphur content in the fuel and the excess air

Optimisation of the steam system is to be done in such a way so as to reach the maximal possible utilisation of the heat contained in the exhaust gas. In this sense the optimisation is reduced to selecting the steam circuit parameters for which the steam turbine reaches the highest power. The area of search for optimal steam circuit parameters is to be narrowed to

**\_\_\_\_\_\_ \_\_\_\_\_\_ \_\_\_\_\_\_ \_\_\_\_\_\_ \_\_\_\_\_\_**

reveal enhanced resistance to acid corrosion.

water temperature, i.e. texh > tFW + (15 – 20oC).

box is within 50 – 60oC.

60

coefficient

**4.2 Optimising the steam circuit** 

70

80

90

100

**Acid dew-point [oC]**

110

120

130

140

150

It is possible to implement a combined system consisting of a Diesel engine as the leading engine, a power gas turbine, and a steam turbine circuit utilising the heat contained in the Diesel engine exhaust gas. Such systems can reveal thermodynamic efficiencies comparable with combined gas turbine circuits connected with steam turbines.

## **5.1 Power range of combined systems**

Depending on the adopted variant and the main engine load, the use of the combined system makes it possible to increase the power of the power plant by 7 to 15 % with respect to the conventional power plant burning the same rate of fuel. Additional power is obtained by the system due to the recovery of the energy contained in the exhaust gas leaving the piston internal combustion engine. Thus the combined system decreases the specific fuel consumption by 6,4 – 12,8 % compared to the conventional power plant.

In the examined systems the power of the steam turbine is higher than that of the power turbine by 6-29 %, depending on the system variant and the main engine load.

## **5.2 Efficiency of combined systems**

The use of the combined system for ship propulsion increases the efficiency of the propulsion system, and decreases the specific fuel consumption. Additionally, it increases the propulsion power without additional fuel consumption.

Like the power, the efficiency of the combined system increases with respect to the conventional power plant by 7 to 15% reaching the level of 53 - 56% for maximal power ranges. These efficiency levels are comparable with the combined systems based on the steam/gas turbines, Fig. 1. For partial loads the efficiency curves of the combined system

Possible Efficiency Increasing of Ship Propulsion and Marine Power Plant with

**Indices:** 

a - air

B - Boiler C - Compressor combi - combined system D - Diesel engine d - supercharging exh - exhaust passage

f - fuel FW - feet water g - exhaust gas inlet - inlet passage

bar - barometric conditions

k - parameters in a condenser o - live steam, calculation point

PT - Power turbine ST - Steam turbine

T - Turbine TC - Turbocharger

**7. References** 

2585

2585

ss - ship living purposes

the System Combined of Marine Diesel Engine, Gas Turbine and Steam Turbine 67


No.1(59), (2009), pp. 47-52, ISSN 1233-2585

INC., ISBN 0-88173-076-9, USA

Dzida, M. (2009). On the possible increasing of efficiency of ship power plant with the

Dzida, M. & Mucharski, J. (2009). On the possible increasing of efficiency of ship power

Dzida, M.; Girtler, J.; Dzida, S. (2009). On the possible increasing of efficiency of ship power

Kehlhofer, R. (1991). *Combined-Cycle Gas & Steam Turbine Power Plants*, The Fairmont Press,

system combined of marine diesel engine, gas turbine and steam turbine at the main engine - steam turbine mode of cooperation. *Polish Maritime Research,* Vol. 16,

plant with the system combined of marine diesel engine, gas turbine and steam turbine in case of main engine cooperation with the gas turbine fed in parallel and the steam turbine. *Polish Maritime Research,* Vol 16, No 2(60), pp. 40-44, ISSN 1233-

plant with the system combined of marine diesel engine, gas turbine and steam turbine in case of main engine cooperation with the gas turbine fed in series and the steam turbine. *Polish Maritime Research,* Vol 16, No 3(61), pp. 26-31, ISSN 1233-

with the Diesel engine are more flat than those for the combined turbine systems (smaller efficiency decrease following the load decrease) .

In the combined system the maximal efficiency is reached using particular system components:


#### **5.3 Ecology**

Along with the thermodynamic profits, having the form of efficiency increase, and the economic gains, reducing the fuel consumption for the same power output of the propulsion system, the use of the combined system brings also ecological profits. A typical newgeneration low-speed piston engine fed with heavy fuel oil with the sulphur content of 3% emits 17g/kWh NOx, 12g/kWh SOx and 600g/kWhCO2 to the atmosphere. The use of the combined system reduces the emission of the noxious substances by, respectively, g/kWh NOx, g/kWh SOx and g/kWhCO2. The emission decreases by % with respect to the standard engine, solely because of the increased system efficiency, without any additional installations.

Depending on the adopted solution, the combined power plant provides opportunities for reaching the assumed power of the propulsion system at a lower load of the main Diesel engine, at the same time also reducing the fuel consumption.

The article presents the thermodynamic analysis of the combined system consisting of the Diesel engine, the power gas turbine, and the steam turbine, without additional technical and economic analysis which will fully justify the application of this type of propulsion systems in power conversion systems.

#### **6. Nomenclature**


## **Indices:**

66 Advances in Gas Turbine Technology

with the Diesel engine are more flat than those for the combined turbine systems (smaller

In the combined system the maximal efficiency is reached using particular system



Along with the thermodynamic profits, having the form of efficiency increase, and the economic gains, reducing the fuel consumption for the same power output of the propulsion system, the use of the combined system brings also ecological profits. A typical newgeneration low-speed piston engine fed with heavy fuel oil with the sulphur content of 3% emits 17g/kWh NOx, 12g/kWh SOx and 600g/kWhCO2 to the atmosphere. The use of the combined system reduces the emission of the noxious substances by, respectively, g/kWh NOx, g/kWh SOx and g/kWhCO2. The emission decreases by % with respect to the standard engine, solely because of the increased system efficiency, without any additional

Depending on the adopted solution, the combined power plant provides opportunities for reaching the assumed power of the propulsion system at a lower load of the main Diesel

The article presents the thermodynamic analysis of the combined system consisting of the Diesel engine, the power gas turbine, and the steam turbine, without additional technical and economic analysis which will fully justify the application of this type of propulsion



engine, at the same time also reducing the fuel consumption.

cg, ca - specific heat of exhaust gas and air, respectively

g, a - isentropic exponent of exhaust gas and air, respectively

systems in power conversion systems.

i - specific enthalpy m - mass flow rate

Wu - calorific value of fuel oil

be - specific fuel oil consumption

efficiency decrease following the load decrease) .

components:

**5.3 Ecology** 

installations.

**6. Nomenclature** 

N - power p - pressure T,t - temperature



## **7. References**


**Part 2** 

**Gas Turbine Systems** 

