1 Introduction 1.1 Marine fuel and emission regulations

Two types of fuel are used by the marine industry as follows: marine distillate, which is also known as Marine Diesel Oil (MDO), and residual fuel oil, which is referred to as bunker fuel, heavy fuel oil, or marine fuel oil. The two fuel types can also be blended together to create a fuel known as intermediate fuel oil. In adddition, marine distillate and low sulfur diesel can be blended to achieve the desired balance between sulfur content and price; low sulfur marine gas oil is an example of this type of blend (Kołwzan and Narewski, 2012; Eyringer et al., 2005).

Global emissions standards currently target a reduction in marine sulfur emissions. Therefore, ships will soon be required to use lower sulfur fuels or emission control devices to reduce the sulfur content to allowable levels. The International Maritime Organization (IMO) controls marine pollution under the MARPOL convention treaty, and NO_x emission limits are set for diesel engines according to engine maximum operating speed (n, r/min) as shown in Table 1. Tier Ⅰ and Tier Ⅱ limits are applicable globally, whereas Tier Ⅲ standards apply only in NO_x Emission Control Areas (ECAs) and will apply to all ships constructed on or after January 2016 that have engines over 130 kW and operate inside an ECA-NO_x area. In addition, the IMO has stipulated that the maximum sulfur content in any fuel used onboard ships should be reduced from 4.5% to 3.5% as of January 2012. Furthermore, it is stated that from January 2020 the sulfur content should not exceed 0.5% (Greensmith, 2010; Herdzik, 2011; Figari et al., 2011).

1.2 Compliance with marine SO_x and NO_x legislations

There is a strong incentive for ship owners and operators to explore the use of alternative fuels and satisfy the lower fuel sulfur and NO_x limits. One approach used to comply with the NO_x and SO_x limits is to firstly use a treatment device known as a scrubber for SO_x reduction and then use selective catalytic reduction equipment to reduce NO_x emissions. However, when considering the use of these technologies it is necessary to consider the cost tradeoff of their installation, operation, and maintenance versus the cost of the alternative fuel (McGill et al., 2013).

O fuels such as natural gas, propane and hydrogen can be considered as alternatives to marine fuel. Not only are these fuels very low in sulfur content, but they also combust; thus, NO_x, PM, and CO₂ emissions are also reduced. Furthermore, they can be carried in compressed or liquefied states. However, these fuels require different fuel handling systems, fuel tanks, and gas burning engines that are currently not in use on most ships ( ElGohary and Seddiek, 2014).It is evident that compliance with the new emission requirements will raise operating costs for ship owners and operators in terms of the construction of new ships with more complicated fuel systems and that some ships will also require treatment devices and the use of more expensive low‐sulfur fuels when in ECAs and other low‐sulfur compliant ports and coastal waters. In addition, existing ships that do not possess dual tanks may have to be retrofitted with dual fossil fuel systems so they can perform fuel switching when they enter an ECA (Brown and Holtbecker, 2007; Lin and Lin, 2006). So, searching for new alternatives for current marine fuels can be considered as a competitive solution for the emission problem as the operating costs will be increased for current ships to be complied with new emission regulations.

1.3 Alternative fuels for marine gas turbines

Natural gas and hydrogen fuels are considered alternative fuels for use in marine gas turbines; their use would reduce exhaust gases emissions in line with international marine regulations. Although natural gas delivers lower emission levels than petroleum products, the problem of limited resources still remains. However, because natural gas is a fossil fuel, it is possible to make it ready for use without the need for further scientific development ( ElGohary, 2012; 2013; ElGohary and Seddiek, 2012). In addition, for decades hydrogen gas has been considered the fuel of the future, and scientific research concerning the use of hydrogen in transportation began shortly after the first oil crisis ( De Luchi, 1989; Welaya et al., 2011; ElGohary et al., 2008; 2014a).

The combustion of gaseous fuels inside internal combustion engines has been the subject of global research programs. Both the following types of internal combustion engines have been studied: the compression ignition engine and the spark ignition engine. In addition, lean burn concepts have been investigated in relation to achieving low emission conditions (Zhang and Frankel, 1998; Huang et al., 2006; Papagiannakis and Hountalas, 2004). However, there are many problems associated with the use of natural gas and hydrogen fuels in these engines, such as engine knocking, and problems with the air fuel ratio and intake temperature (Seddiek et al., 2015; Hailin and Ghazi, 2004; White et al., 2006; Gan et al., 2011). Therefore, to optimize the injection timing inside the engine cylinders and the cylinder geometry, it is necessary to enable accurate control of parameters to avoid engine knocking and high emission formation levels. Another application for gaseous fuels is the gas turbine engine. From the emissions viewpoint, natural gas produces fewer emissions in gas turbines than in internal combustion engines except for CO₂ because of the higher fuel consumption in gas turbines (Chicco et al., 2008; ElGohary et al., 2014b). In addition, hydrogen produces fewer emissions in almost all engine operating conditions because of the higher combustion temperature in hydrogen engines, although higher NO_x rates have been found in some cases.

Solutions for reducing NO_x emissions include exhaust gas recirculation and catalytic reduction filters (Heffel, 2003; Ho et al., 2008). There are basically three methods that can be applied in a gas turbine powered by petroleum products or natural gas to reduce emissions as follows: dry low NO_x combustors, flame dilution by the addition of steam, or adding catalytic reducers to the exhaust system. However, not all of these techniques can be used with hydrogen due to its special combustion characteristics (Ma et al., 2003; Chiesa et al., 2005).

There are two main problems of using hydrogen in the marine field as follows: hydrogen storage and hydrogen production cost. The best option for hydrogen storage is in a liquified phase with a density of 70.1 kg/m³, which is a very small value compared with ordinary liquid fuels with densities ranging from 850 to 1 010 kg/m³ (Woud and Stapersma, 2002). However, the main problem in storing liquid hydrogen is the weight of the system involved; cryogenic tanks used to store liquid hydrogen are heavily insulated to prevent evaporation of the −253°C stored liquid fuel, and this insulation can increase the tank’s weight to between two and four times that of an equivalent diesel tank with the same energy content (Veldhuis et al., 2005).

2 LM2500 marine gas turbine model

The LM2500 gas turbine is categorized as an aero-derivative gas turbine (derived from a previous model used on aircraft). This engine is a turbo shaft gas turbine, which is suitable for marine use as well as land based electric generation plants. In 2004, the number of installed LM2500s reached more than 1000 units. The turbine is mainly used in naval applications but has recently started to appear in the commercial market, particularly in the passenger vessel field; the famous ocean liner Queen Mary Ⅱ has two units installed for electrical generation purposes, and fast ferries use the LM2500 for propulsion combined with diesel engines in CODOG and CODAG configurations. Table 2 presents the main characteristics of the model.

For marine gas turbine engines, a separate free power turbine is introduced to separate the compressor turbine from the torque variations occurring because of variations in propeller loading (Harrington, 1992). Thus, a new point is introduced in the cycle to represent air conditioning between the two turbines. Fig. 1 shows a schematic representation of the cycle with a free power turbine, and Fig. 2 shows a gas turbine cycle with a free power turbine installed after the compressor turbine wherein point 4 represents the exhaust gases from the first turbine and point 5 represents the exhaust point of the gas turbine.

3 Gas turbine cycle description

The ISO design conditions of the compression process (Harrington, 1992) state that the initial temperature and pressure of the air should be 15°C and 1.0 bar, respectively. The compression process is first assumed to be isentropic in reaching point (2^*) (shown in Fig. 1), and to reach point (2), the point is corrected using isentropic efficiency after the compressor. All calculations are made using air enthalpy rather than air temperature to accommodate the effect of varying air specific heat at different temperatures during the cycle. Enthalpy at any point (i) with temperature of T_i and specific heat at constant pressure at this temperature Cp_i can be calculated using:

${h_i} = C{p_i} \times {T_i}$

(1)

The isentropic air enthalpy after the compressor (h₂^*) is calculated at compression pressure (P₂) at a constant entropy as point (1) at the compressor inlet as follows:

${\eta _C} = \frac{{h_2^* - {h_1}}}{{{h_2} - {h_1}}}$

(2)

After the compression process, the air enters the combustor at a high enough pressure and temperature to continue the fuel burning process occurring inside. However, the amount of heat added to the air from the fuel cannot be determined unless the cycle efficiency and output power are known. It is more convenient to express this amount using the unit of air mass flow, and this is obtained by dividing the fuel calorific value (CV) by the total air-fuel ratio (AF_T), which is the product of the fuel stoichiometric ratio (AF) and the excess air factor (λ); thus, the fuel calorific value is for unit air mass flow rather than unit fuel mass flow as follows:

${\text{C}}{{\text{V}}_{{\text{air}}}} = \frac{{{\text{CV}}}}{{{\text{A}}{{\text{F}}_{\text{T}}}}}$

(3)

${\text{A}}{{\text{F}}_{\text{T}}} = {\text{AF}} \times \lambda $

(4)

This amount (CV_air) is added to the air enthalpy after the compressor in the combustor to reach the air condition before the turbine at point (3) as follows:

${h_3} = {\eta _{{\text{comb}}}} \times {\text{C}}{{\text{V}}_{{\text{air}}}} + {h_2}$

(5)

In the above equation, the combustion efficiency, (η_comb), is used to express the real amount of heat added to the flow inside the combustion chamber; this value is 2% smaller than the amount of heat obtained from combustion of the fuel flow because of the incompletely burnt amount of fuel. It is important to note that the air pressure inside the combustor is not constant and that a small pressure drop takes place due to air friction against the combustor walls; this pressure drop is taken to be 5% of the combustor inlet pressure (P₂) as follows:

${P_3} = {P_2} - \Delta P$

(6)

The air now enters the compressor turbine, which produces enough work for the compressor to continue running as follows:

${W_C} = {W_{{\text{CT}}}}$

(7)

${W_{{\text{CT}}}} = {h_2} - {h_1}$

(8)

Knowing that the work of the compressor turbine is the difference in the air enthalpy across the turbine, the air condition at point (4) after the compressor turbine can thus be calculated as follows:

${W_{{\text{CT}}}} = {h_3} - {h_4} = {h_2} - {h_1}$

(9)

To calculate the air pressure between the turbines at point 4, the isentropic expansion line (3-4^*) must be considered to calculate the isentropic enthalpy of air after the first turbine using the turbine isentropic efficiency as follows:

${\eta _{\text{T}}} = \frac{{{h_3} - {h_4}}}{{{h_3} - h_4^*}}$

(10)

Knowing that (h₄^*) occurs at constant entropy as point (3), the pressure at point (4) can now be determined since two properties are now defined at this point; the enthalpy and the entropy.

For the expansion process in the power turbine, the turbine exhaust pressure is assumed to be the same of the atmospheric pressure as follows:

${P_5} = {P_1}$

(11)

The isentropic enthalpy at the exit of the power turbine, (h₅^*), is firstly determined at atmospheric pressure and constant entropy as point (4). The enthalpy at point (5) is then determined using the turbine isentropic efficiency. Note that the two turbines are assumed to have the same isentropic efficiency as follows:

${\eta _T} = \frac{{{h_4} - {h_5}}}{{{h_4} - h_5^*}}$

(12)

The work of the free power turbine can now be determined as the difference in air enthalpy across the turbine as follows:

${W_{{\text{PT}}}} = {h_4} - {h_5}$

(13)

After determining the air condition at five cycle points, the cycle performance is assessed by calculating the cycle efficiency, the peak temperature ratio, the work ratio, and the mass flow rate of air, fuel, and exhaust. To determine the different flow rates, the output power is obtained according to the power-temperature curve of the gas turbine.

The cycle efficiency (η_cycle) is the ratio between the output work and the input energy to the engine. This can be expressed in specific form using the specific work of the power turbine (W_PT) and the energy content in the fuel (CV_air) as follows:

${\eta _{{\text{cycle}}}} = \frac{{{W_{{\text{PT}}}}}}{{{\text{C}}{{\text{V}}_{{\text{air}}}}}}$

(14)

The peak temperatures ratio (ε) is the ratio between the maximum and minimum temperatures occurring in the cycle, namely the inlet air temperature (T₁) and the combustion temperature (T₃) as follows:

$\varepsilon = \frac{{{T_3}}}{{{T_1}}}$

(15)

The work ratio is the ratio between the work of the power turbine and the total work generated from both turbines and is then determined as follows:

${\text{WR}} = \frac{{{W_{{\text{PT}}}}}}{{{W_{{\text{PT}}}} + {W_{{\text{CT}}}}}}$

(16)

This value measures the effectiveness of the engine and how much power is useful from the sum of power generated inside the engine.

The flow inside the turbines is composed of both air and fuel, but as the study is conducted using air as the only working fluid, the flow rate inside the turbine is considered to be composed of two air flows; the air from the compressor and the injected fuel as follows:

${\dot m_{{\text{air}}}}\left ( {1 + \frac{1}{{A{F_{\text{T}}}}}} \right) = \frac{{{\text{Power}}}}{{{W_{{\text{PT}}}}}}$

(17)

The fuel mass flow rate is then determined according to the air to fuel ratio as follows:

${\dot m_{{\text{fuel}}}} = \frac{{{{\dot m}_{{\text{air}}}}}}{{A{F_{\text{T}}}}}$

(18)

The specific fuel consumption is determined by dividing the mass flow rate by the output power as follows:

$SFC = \frac{{{{\dot m}_f}}}{{Power}}$

(19)

and the exhaust mass flow rate is the sum of both flow rates of the air and the fuel as follows:

${\dot m_{ex}} = {\dot m_{{\text{air}}}} + {\dot m_{{\text{fuel}}}}$

(20)

In this study, the above thermodynamic model is used in an assessment with natural gas in one case and hydrogen in another to enable a comparison with the original case using diesel fuel. Table 3 shows the principle properties of the various fuel types.

4 Comparison of gas turbine performance

parameters using alternative fuels A comparison of gas turbine performance using diesel, natural gas, and hydrogen fuels was made assuming constant power output in both the cases of natural gas and hydrogen (the performance of the engines was assessed based on the same amount of achieved power). The software used in the comparison was the Engineering Equations Solver, where the thermodynamic properties of the substances under study could easily be obtained using the built-in functions and data. The program is used to show the effect of the fuel properties on the gas turbine performance, assuming initial conditions of inlet temperature and pressure of 15°C and 1 bar, respectively. The temperatures used in study are in the range between −25°C and 55°C, which correspond to normal operating conditions for ships worldwide.

4.1 Cycle efficiency

The gas turbine performance is critically limited by the predominating ambient temperature, and this occurs mainly in hot and dry regions; the power output is inversely proportional to the ambient temperature. In addition, the drop in inlet air temperature for gas turbine provides an increase in air density, and consequently, the air mass flow rate is elevated. This behavior increases the power output and efficiency for the gas turbine. This study determined that both gaseous fuels provide a lower efficiency than the original case of diesel for the same power output as shown in Fig. 3; the reason for this is clarified later in the study when air flow rates for the three cases are discussed.

4.2 Specific fuel consumption

The higher calorific value of both natural gas and hydrogen compared to that of diesel reduces the quantity of fuel required to give the same heat output. It was thus observed that the specific fuel consumption for natural gas and hydrogen is lower than that of diesel. Although this could represent an advantage with respect to available storage space onboard ships, the advantage could not be realised because of the lower density of these two fuels compared to diesel. Figs. 4 and 5 show the specific fuel consumption comparison of gas turbines as a function of inlet air temperature and compression ratios for diesel, natural gas, and hydrogen fuels. It is shown that the specific fuel consumption of the three fuels increases with an increase in the inlet air temperature, and conversely, the fuel consumption rates decrease with an increase in the compression ratio.

4.3 Peak temperature

In this study, the different parameters of the thermodynamic model in both cases of hydrogen and natural gas were controlled to avoid obtaining higher combustion temperatures than in the case of diesel with the aim of avoiding an increase in turbine blade cooling. Therefore, the peak temperatures obtained for the natural gas and hydrogen are lower than that obtained for diesel as shown in Figs. 6 and 7.

4.4 Exhaust temperature

Due to the lower maximum temperatures achieved in the case of both gaseous fuels, the exhaust temperatures followed the same trend, with lower temperatures obtained for hydrogen and natural gas. The main parameters affecting gas turbine exhaust temperatures are inlet air temperature and the compression ratio. The temperature of exhaust gases increases with an increase in the inlet air temperature (as shown in Fig. 8); in contrast, the exhaust gas temperature decreases as the compression ratio increases (as shown in Fig. 9) in relation to increased rates of inlet air mass flow.

4.5 Work ratio

The work ratio is the ratio between the useful work developed inside the engine and the total work developed. Due to the lower efficiency of gaseous fuels, the work ratio of the fuels also appears inferior to that of diesel, as shown in Figs. 10 and 11. The work ratio decreases in line with an increase in the compression ratio for the three fuels. This occurs because the compressor inlet temperature and the intake air density dictate the mechanical work required by the compression process and the quantity of fuel used in obtaining the necessary temperature at the gas turbine inlet.

4.6 Air mass flow rate

The gas turbine inlet air mass flow rate affects the gas cycle efficiency, and the gas turbine output power is directly proportional and limited by the air mass flow rate. As the compressor has a fixed capacity for a given rotational speed and volumetric flow rate of air, the volumetric capacity remains constant, and thus there is a variation in the mass flow rate of air as it enters the gas turbine, which is dependent on its specific mass. Therefore, the mass flow rate of air is dependent on the temperature and relative humidity of the ambient air, and the mass flow rate of air through the engine dictates engine performance. In addition, any restrictions against the smooth flow of air through the engine limit the performance of the engine. The pressure ratio of the compressor, the engine operating temperatures, and the individual component efficiencies also influence both the performance and the efficiency of the overall engine. All such factors need to be considered during the design of the gas turbine, and an optimum pressure ratio, turbine inlet temperature, and air mass flow rate need to be selected to obtain the required performance in the most efficient manner. In addition, individual engine components need to be designed to minimize flow losses with the aim of maximizing component efficiencies.

The air flow rates for natural gas and hydrogen fuels are higher than that of diesel, as shown in Figure 12, because of the higher stoichiometric air to fuel ratio for gaseous fuels than that of diesel fuel. In addition, the excess air factors used for the gaseous fuels are higher than that of diesel, and this introduces another increase in the amount of air entering the engine. This is the main reason for the lower efficiency of gaseous fuel engines; a greater amount of air is used to keep the temperature in a reasonable range, and thus part of the heat generated by fuel combustion is lost to enable a decrease in maximum temperature. Therefore, it would potentially be necessary to use a bigger compressor to achieve an identical performance for all three fuel cases.

4.7 Exhaust mass flow rate

Following the increase in air mass flow rates for natural gas and hydrogen, the exhaust mass flow rate is also increased, although by a smaller percentage compared to the diesel case, as shown in Fig. 13. This smaller increase is due to the lower fuel flow rates for both the gaseous fuels. The change in mass flow rates through the compressor and turbines leads to a deviation in compressor turbine matching; therefore it is not possible for gaseous fuels to replace ordinary fuels without incurring large changes in performance.

At a constant gas turbine speed, the compressor pumps a constant volume of air into the engine without consideration of air mass or density. If the density of air decreases, the same volume of air will have a smaller mass, and less exhaust gas will be produced. Fig. 14 shows the exhaust gas mass flow rate as a function of the compression ratio; as the compression ratio increases, the exhaust gas mass flow rate decreases.

This study shown that gaseous fuels deliver a good performance compared to diesel fuel, and that natural gas is the currently the best choice of hydrocarbon to replace diesel fuel as it can be supplied at a relatively low price and is highly availabile. Clean combustion assists with increasing maintenance intervals for gas turbine components. However, fossil fuel reserves will ultimately be exhausted, particularly with the increasing global energy demand . Thus, hydrogen could eventually be introduced to replace fossil fuels once the problems associated with its application are solved, such as the cost of its production and its storage onboard vessels. The application of diesel fuel in gas turbines is not preferred due to the associated high emission levels in comparison with gaseous fuels, and the increasing price, which is considered a problem due to the high fuel consumption rates of gas turbines.

5 Conclusions

A thermodynamic analysis of a natural gas and hydrogen-fueled marine gas turbine was presented in relation to the proposed use of these fuels as alternatives to satisfy the requirements of international marine regulations. A thermodynamic analysis of the LM2500 marine gas turbine resulted in the following conclusions:

Firstly, liquid fuels, which are currently used with marine gas turbines, are associated with a number of environmental and economic issues, whereas natural gas and hydrogen fuels represent a positive solution for overcoming the difficulties associated with current marine liquid fuels.

Secondly, the thermodynamic performance of natural gas fueld is found to be close to that of diesel oil, and its maximum cycle temperature is 1474 K, which is close to that of diesel fuel (1485 K). With an efficiency reduction of about 0.25% using ISO design conditions, natural gas is thus considered to be an excellent replacement for diesel fuel.

Thirdly, it was determined that for hydrogen fuel a number of modifications would be required to enable optimum performance. The gas turbine thermal efficiency was found to be 1% less in the case of hydrogen compared to diesel, with a maximum cycle temperature of 1445 k. It is thus considered that the application of hydrogen in gas turbines cannot be contemplated as a near-future solution until the problems associated with its use are solved using proper techniques that are technically and economically feasible.

Finally, gaseous fuels deliver good performances compared to diesel fuel, but to achieve such performances the engine compressor and turbines need to be modified to accommodate differing flow rates, particularly in the case of hydrogen, where large differences in flow rates were observed. Therefore, hydrogen fuel could be considered as an alternative new fuel for marine gas turbines in the long term, but in the short term natural gas represents a positive solution for marine applications.

Nomenclatures

AF	Fuel stoichiometric ratio
AFT	Total air-fuel ratio
Cp	Specific heat at constant pressure, kJ/kg K
CVair	Energy content in fuel, kJ/kg
CV	Fuel calorific value, kJ/kg
H	Enthalpy, kJ⁄kg
LCV	Lower calorific heat value, kJ/kg
$\dot m$	Mass flow rate, kg/s
$\dot m$air	Air mass flow rate, kg/s
$\dot m$Ex	Exhaust mass flow rate, kg/s
$\dot m$fuel	Fuel mass flow rate, kg/s
p	Pressure, bar
R	Compression ratio
SFC	Specific fuel consumption, kg/kW hr
T	Temperature, K
Ta	Air Ambient temperature, K
T3	Maximum cycle temperature, K
Wc	Compressor specific work, kJ/kg
WCT	Work of turbine compressor, kJ/kg
WPT	Specific work of power turbine, kJ/kg
WPT	Work of power turbine, kJ/kg
WR	Work ratio
WT	Turbine work

Greek letters

ε	Maximum temperature ratio
γair	Air specific heat ratio
γgas	Gas specific heat ratio
η	Efficiency
ηcomb.	Combustor efficiency
ηcycle	Cycle efficiency
ηT	Turbine isentropic efficiency
λ	Excess air factor

Abbreviations

CO2	Carbon dioxide
CO	Carbon monoxide
CODAG	Combined dieseland gas turbine
CODOG	Combined diesel or gas turbine
ECA	Emission control area
HFO	Heavy fuel oil
IMO	International Maritime Organization
MARPOL	International marine pollution prevention
	convention
NOx	Nitrogen oxides emissions
PM	Particulate matter
SOx	Sulfur oxide emissions