energies

Article Comparison of Saturated and Superheated Steam Plants for Waste-Heat Recovery of Dual-Fuel Marine Engines

Marco Altosole 1, Giovanni Benvenuto 2,*, Raphael Zaccone 2 and Ugo Campora 3 1 Dipartimento di Ingegneria Industriale (DII), School of Polytechnic and Basic Sciences, University of Naples “Federico II”, 80138 Naples, Italy; [email protected] 2 Dipartimento di Ingegneria Navale, Elettrica, Elettronica e delle Telecomunicazioni (DITEN), University of Genoa, Via Montallegro 1, I-16145 Genova, Italy; [email protected] 3 Dipartimento di Ingegneria Meccanica, Energetica, Gestionale, Trasporti (DIME), University of Genoa, Via Montallegro 1, I-16145 Genova, Italy; [email protected] * Correspondence: [email protected]; Tel.: +39-348-665-3079



Received: 16 January 2020; Accepted: 18 February 2020; Published: 22 February 2020


Abstract: From the working data of a dual-fuel marine engine, in this paper, we optimized and compared two waste-heat-recovery single-pressure steam plants—the first characterized by a saturated-steam Rankine cycle, the other by a superheated-steam cycle–using suitably developed simulation models. The objective was to improve the recovered heat from the considered engine, running with both heavy fuel oil and natural gas. The comparison was carried out on the basis of energetic and exergetic considerations, concerning various aspects such as the thermodynamic performance of the heat-recovery steam generator and the efficiency of the Rankine cycle and of the combined dual-fuel-engine–waste-heat-recovery plant. Other important issues were also considered in the comparison, particularly the dimensions and weights of the steam generator as a whole and of its components (economizer, evaporator, superheater) in relation to the exchanged thermal powers. We present the comparison results for different engine working conditions and fuel typology (heavy fuel oil or natural gas).

Keywords: heat-recovery systems; marine propulsion plants; combined-cycle power plants; emission reduction

1. Introduction Currently, most engines used for ship propulsion still consist of diesel engines fueled by traditional fuel types, such as Heavy Fuel Oil (HFO) and Marine Diesel Oil (MDO). However, in recent years, there has been an increasingly widespread use of Dual-Fuel (DF) engines capable of burning both diesel fuels and natural gas (NG), motivated by a series of technical and environmental advantages. The growing demand for a reduction in the polluting emissions of ships, promoted by increasingly restrictive international regulations, has in fact favored interest in engines powered by natural gas that could significantly reduce emissions of sulfur oxides, nitrogen oxides, particulate matter, and carbon dioxide [1]. Regulations adopted by the International Maritime Organization (IMO), initially focused on sulfur and nitrogen oxides [2], were subsequently oriented toward carbon dioxide emissions [3,4] due to the need of reducing the greenhouse effect. The reduction of CO2 emissions can be achieved by using fuels characterized by low carbon content, such as natural gas [5], but also by improving the efficiency of marine propulsion systems, with positive effects on the reduction of fuel costs. An example of a typical reduction of pollutant emissions of an NG-fueled engine in comparison with a diesel engine fueled with HFO was reported in [6]: 25 30% carbon dioxide, 25% carbon monoxide, ÷

Energies 2020, 13, 985; doi:10.3390/en13040985 www.mdpi.com/journal/energies Energies 2020, 13, 985 2 of 21

85% NOx, 98 100% SOx, 98 99% particulate matter. With regard to the other aspect mentioned above, ÷ ÷ namely the efficiency improvement of ship propulsion plants, with positive implications on emissions and fuel costs, many solutions have recently been proposed, including those that are based on the recovery of heat contained in the exhaust gases of the main engine by means of Waste Heat Recovery (WHR) systems. Among these solutions we can mention turbocharging [7,8], turbocompounding [9], thermochemical recuperation [10], thermoelectrics [11], cabin cooling [12], and of course WHR steam plants [13]. The adoption of this last solution, of particular interest in the marine propulsion sector, allows to satisfy the vessel heat necessity and, at the same time, to produce electric energy by means of a steam turbine, then reducing the power of on-board diesel generators and their emissions. It should also be noted that the use of steam plants in WHR systems is not limited to diesel engines. In [13], possible applications were examined in ships using gas turbines as a prime mover, where the possibility of exploiting part of the thermal energy of exhaust gases by means of a steam plant, thus generating a Combined Gas and Steam (COGAS) system, makes it possible to increase the efficiency of the propulsion system and limit polluting emissions. More generally, with regard to the use of steam plants in recent and current applications to marine propulsion, the case of commercial Liquefied Natural Gas (LNG) carriers is particularly significant. The choice of propulsion system for this type of ship depends on the type of transported cargo, i.e., liquefied natural gas at a temperature of approximately 160 C, below its boiling point at the corresponding pressure inside the tanks. Despite the thermal − ◦ insulation of the tanks, designed to minimize heat supply from the outside, even a small part of this heat causes partial evaporation of the load, known as boil-off-gas (BOG). Depending on the modality of using the BOG and the type of power transmission (mechanical or electric drive), it is possible to identify main propulsion solutions for LNG carriers. In the case of BOG used as fuel, the solution adopted since the 1960s is that of a steam turbine with mechanical drive. This solution is still the most common on the market today [14], with a percentage of the LNG fleet above 40% (44% in May 2017). However, since 2004, several other solutions have been proposed with the use of both gas turbines and diesel engines. Among these are DF gas turbines with an electric drive, Combined Gas turbines and Electric Steam turbines (COGES), DF medium-speed diesel engines with an electric drive, and DF low-speed diesel engines with a mechanical drive. In the case of recovered BOG, the most considered solution by the market after 2010 [14] is that of the single-fuel, two-stroke, low-speed engine with a mechanical drive and a re-liquefication plant, which has currently reached a percentage of the LNG fleet of about 40% (41% in May 2017). More details on propulsion systems for LNG carriers and their evaluation can be found in [14,15].

2. WHR Steam Plants Overview Many examples of WHR systems from marine diesel engines were presented in the literature, some of which are mentioned in references at the end of this paper [16–28]. Of the last two papers [27,28], the former is on the effects of the deterioration of steam-plant components on the performance of WHR systems, and the latter presents an interesting comparison of an organic and a steam Rankine cycle applied to WHR plants. To the authors’ knowledge, only [6,26] refer to DF marine engines combined with WHR steam plants. All cited papers refer to single- and/or dual-pressure WHR superheated steam plants, with the only exception of [26], where a single-pressure saturated steam plant was used for waste-heat recovery of the main engine. In a previous paper of the authors [29], six WHR plants of different types (single- or dual-pressure, saturated- or superheated-steam) were optimized and compared for application to the propulsion engines of an existing cruise ferry. We found that, mainly in the case of single-pressure WHR steam plants, there was no substantial difference between the state of used steam (saturated or overheated) with regard to the power of the steam turbine and the efficiency of the Combined Engine–WHR system. Energies 2020, 13, 985 3 of 21

The main objective of this article is to demonstrate that the prevailing opinion from bibliographic analysis, which considers the use of the superheater in WHR systems convenient, has no scientific basis. For this reason, we wanted to contribute to clarify this issue by using a simulation methodology of these systems previously developed and validated. Following this observation, we carried out a thorough comparison between two single-pressure WHR plants, one characterized by saturated steam, the other by superheated steam, applied to the same DF four-stroke marine engine. Both WHR steam plants, characterized by a similar layout, were optimized by means of a simulation code, with a procedure presented by the authors in [24], aimed at achieving the best efficiency of the combined engine–WHR system with the engine running at Normal Continuous Rating (NCR). The WHR steam plants were then compared on the basis of energetic and exergetic considerations, with regard to the thermodynamic performance of the Heat Recovery Steam Generator (HRSG), the dimensions and weights of the steam generator as a whole and of its components (economizer, evaporator and superheater) in relation to the exchanged thermal powers, the Rankine cycle characteristics, and the efficiency of the combined DF engine–WHR plant. The obtained results from the comparison for different engine working conditions and fuel type (HFO or NG) are then discussed and presented.

3. Case Study In order to define the schemes of the WHR steam plants to be compared, it is necessary to know the characteristics of the propulsion engine to which these plants are applied and the electric- and thermal-power needs of the considered vessel. The vessel in question is an existing cruise ferry of the GNV shipping line (overall length = 211 m; displacement = 49,257 tons; cruising speed = 22 knots). The propulsion system consisted of two controllable pitch propellers, each driven by two Wärtsilä 16V46C four-stroke diesel engines, of which the Maximum Continuous Rating (MCR) power is 16,800 kW at 500 rpm. In normal sea going conditions, only one of the two engines for each propeller shaft is active, and runs at Normal Continuous Rating (NCR), about 76% of the MCR. Electric users of the cruise ferry are powered by four Wärtsilä 6R32LNE diesel generators (2430 kWe at 720 rpm). The required thermal power by the Ship Steam Services (SSS) is satisfied by 1.15 kg/s of saturated steam at a pressure of at least 7 bar. On the basis of the shipowner indications, we considered that only part of the steam produced by a single WHR steam plant (one for each engine) was used to satisfy the thermal-power demand of the ship, while the remaining part was sent to a steam turbine for the production of electric energy. In the simulation, the originally adopted diesel engine was substituted by a MAN 51/60 18V DF four-stroke engine, built by Man Diesel & Turbo, characterized by 17,550 kW of MCR power at 500 rpm [30]. A comparison of the main data of the two engines is shown in Table1. The choice of the new engine was motivated by the greater power and efficiency of the selected engine, as well as by more complete information availability. To satisfy the ship’s normal seagoing propulsion power and half the required SSS thermal power (0.57 kg/s), this new engine works at 488 rpm and at 75% MCR load, corresponding to 13,162.5 kW (NCR). When the engine is powered by natural gas (NG), the HFO fuel-tank heating is not necessary; consequently, ship SSS thermal power is reduced by 70% (only 0.17 kg/s of saturated steam for each running engine, at a pressure of at least 2 bar, sufficient in this condition). The manufacturer data of the new DF engine [30], reported in Figure1, show the remarkable difference in (a) exhaust-gas temperature, (b) mass-flow rate, and (c) thermal power between the two possible engine fuels (HFO or NG). Energies 2019, 12, x FOR PEER REVIEW 4 of 20

To satisfy the ship’s normal seagoing propulsion power and half the required SSS thermal power (0.57 kg/s), this new engine works at 488 rpm and at 75% MCR load, corresponding to 13,162.5 kW (NCR). When the engine is powered by natural gas (NG), the HFO fuel-tank heating is not necessary; consequently, ship SSS thermal power is reduced by 70% (only 0.17 kg/s of saturated steam for each running engine, at a pressure of at least 2 bar, sufficient in this condition). The manufacturer data of the new DF engine [30], reported in Figure 1, show the remarkable difference in (a) exhaust-gas temperature, (b) mass-flow rate, and (c) thermal power between the two Energies 2020, 13, 985 4 of 21 possible engine fuels (HFO or NG).

400 40 12 HFO HFO HFO NG NG 11 NG 375 35 10 350 30 9 325 25 8

300 20 7 Exhaust Gas Power [MW] Power Gas Exhaust Exhaust gas temperature [°C] temperature gas Exhaust 275 [kg/s] flow mass gas Exhaust 15 6 50 60 70 80 90 100 50 60 70 80 90 100 50 60 70 80 90 100 Engine load [%] MCR Engine load [%] MCR Engine Load [%] MCR (a) (b) (c) Figure 1. (a) Engine exhaust gas temperature, (b) mass-flow rate, and (c) thermal power versus engine Figure 1. (a) Engine exhaust gas temperature, (b) mass-flow rate, and (c) thermal power versus engine load for two possible engine fuels. load for two possible engine fuels.

4. Schemes of the SelectedTable WHR 1. Comparison Steam Plants of main data of current and new engine.

In this section, the two considered WHRCurrent steam Engineplants, applied to the Newabove-mentioned Engine MAN Engine Parameters dual-fuel engine, are briefly described, with referenceWärtsilä 16V to the 46C schemes shown MAN in 51 Figures/60 18V DF2 and 3. Length (mm) 12,871 13,644 Height (mm) 5854 5517 Width (mm) 4678 4713 Dry weight (tons) 233 265 Cylinders number 16 18 Bore (mm) 460 510 Stroke (mm) 580 600 Fuel type HFO HFO/NG MCR power (kW) at 500 rpm 16,800 17,550 B.m.e.p. (bar) 26.14 19.09 Injection type single double Efficiency (%) 45.97 47.68 (NG)

4. Schemes of the Selected WHR Steam Plants In this section, the two considered WHR steam plants, applied to the above-mentioned MAN dual-fuelFigure engine, 2. Waste-heat-recovery are briefly described, (WHR) withsingle-pressure reference saturated-steam to the schemes plant shown (T1g in = FiguresT2g). Abbreviations2 and3. ; TheAC: aftercooler; first was a AF: single-pressure air filter; C: turbocharger saturated-steam compressor; plant, CP: of evap whichorator the circulating scheme is pump; shown DF in ENG: Figure 2, also considereddual fuel engine; in a previousE: economizer; paper EG: of theelectric authors generator; [26]. EV: This evapor WHRator; steam HWT: plant heat is water basically tank; composed ILV: of anisenthalpic HRSG comprising lamination an valve; economizer JW: jacket (E) water; and MFP: an evaporator main feed (EV)pump; that SC: feeds scavenging a steam system; turbine SCai: (ST) to drivescavenging an electric air generator inlet; SCao: (EG). scavenging air outlet; SCO: steam condenser; SCP: steam cond. pump; SD: Thesteam turbine drum; exhaustSSC: ship steam steam isservices condensed condenses; in a condenser ST: steam turbine; (SCO), fromT: turbocharger which it is turbine; extracted 1g: by a condensingHRSG gas pump inlet; (SCP) 2g: evaporator and preheated gas inlet; in 3g: the evapor Jacketator Water gas (JW)outlet; heat 4g: HRSG exchanger gas outlet; before 0JW: being jacket sent to the Heatwater Water outlet; Tank 0s: heat (HWT), water from tank which outlet; it0sc: is takenmain feed by the pump main outlet; feed pump0sCOND: (MFP). steam cond. pump outlet; 00sc: scavenger water outlet; 1s; economizer outlet; 2s: evaporator circ. pump inlet; 2”s: After being warmed by the hot air leaving the turbocharger compressor in the scavenge air evaporator outlet; 3s: steam turbine inlet; 3sd: SD saturated steam outlet; 4s: steam turbine outlet. preheater (SC), the water is sent to the HRSG economizer (E). To satisfy SSS, part of the saturated steam is taken from the HRSG steam drum (SD), as shown in Figure2.

The second WHR plant is a single-pressure superheated-steam plant, of which the scheme is shown in Figure3. The basic layout of the superheated-steam system was the same as that adopted by the saturated-steam system, but with the addition of a superheater (SH in Figure3) in the HRSG that takes saturated steam from the SD and feeds the steam turbine with superheated steam. Energies 2019, 12, x FOR PEER REVIEW 4 of 20

To satisfy the ship’s normal seagoing propulsion power and half the required SSS thermal power (0.57 kg/s), this new engine works at 488 rpm and at 75% MCR load, corresponding to 13,162.5 kW (NCR). When the engine is powered by natural gas (NG), the HFO fuel-tank heating is not necessary; consequently, ship SSS thermal power is reduced by 70% (only 0.17 kg/s of saturated steam for each running engine, at a pressure of at least 2 bar, sufficient in this condition). The manufacturer data of the new DF engine [30], reported in Figure 1, show the remarkable difference in (a) exhaust-gas temperature, (b) mass-flow rate, and (c) thermal power between the two possible engine fuels (HFO or NG).

400 40 12 HFO HFO HFO NG NG 11 NG 375 35 10 350 30 9 325 25 8

300 20 7 Exhaust Gas Power [MW] Power Gas Exhaust Exhaust gas temperature [°C] temperature gas Exhaust 275 [kg/s] flow mass gas Exhaust 15 6 50 60 70 80 90 100 50 60 70 80 90 100 50 60 70 80 90 100 Engine load [%] MCR Engine load [%] MCR Engine Load [%] MCR (a) (b) (c) Figure 1. (a) Engine exhaust gas temperature, (b) mass-flow rate, and (c) thermal power versus engine load for two possible engine fuels.

4. Schemes of the Selected WHR Steam Plants EnergiesIn2020 this, 13 section,, 985 the two considered WHR steam plants, applied to the above-mentioned MAN5 of 21 dual-fuel engine, are briefly described, with reference to the schemes shown in Figures 2 and 3.

Energies 2019, 12, x FOR PEER REVIEW 5 of 20

The first was a single-pressure saturated-steam plant, of which the scheme is shown in Figure 2, also considered in a previous paper of the authors [26]. This WHR steam plant is basically composed of anFigureFigure HRSG 2.2. comprising Waste-heat-recovery an economizer (WHR)(WHR) single-pressuresingle-pressure(E) and an evaporator saturated-steam saturated-steam (EV) that plant plant feeds (T (T1g1g = a= TsteamT2g2g). AbbreviationsAbbreviations turbine (ST); to driveAC:AC: an aftercooler;electricaftercooler; generator AF:AF: air filter;filter; (EG). C: turbocharger compressor; CP: evaporatorevaporator circulating pump; DF ENG: dualThedual turbine fuel engine; exhaust E: economizer; steam is condensedEG: electric generator; in a cond EV:enser evapor evaporator; (SCO),ator; from HWT: which heat water it is extracted tank; ILV: ILV: by a condensingisenthalpicisenthalpic pump laminationlamination (SCP) and valve; preheated JW: jacketjacket in water;water; the Jacket MFP: mainWater feed (JW) pump; heat SC:exchanger scavenging before system; being SCai: sent to the Heatscavengingscavenging Water airairTank inlet;inlet; (HWT), SCao:SCao: scavengingfrom which air it outlet; is taken SCO: by steam the main condenser; feed pump SCP: steam (MFP). cond. pump; SD: steamAftersteam drum;beingdrum; SSC: warmedSSC: ship ship steam bysteam the services serviceshot air condenses; condenses;leaving ST: the steamST: turbocharger steam turbine; turbine; T: turbochargercompressor T: turbocharger turbine; in the turbine; 1g:scavenge HRSG 1g: air preheatergasHRSG inlet; (SC), gas 2g: inlet; the evaporator water2g: evaporator is gas sent inlet; togas 3g:the inlet; evaporatorHRSG 3g: evapor econom gasator outlet;izer gas (E).outlet; 4g: HRSGTo 4g: satisfy HRSG gas outlet; SSS, gas outlet;part 0JW: of jacket 0JW: the waterjacketsaturated steamoutlet;water is taken 0s:outlet; heat from 0s: water theheat tankHRSG water outlet; steamtank 0sc: outlet; drum main 0sc: feed(SD), main pump as feedshown outlet; pump in 0sCOND: Figure outlet; steam2.0sCOND: cond. steam pump cond. outlet; pump 00sc: scavengerTheoutlet; second 00sc: water scavengerWHR outlet; plant 1s;water economizeris a outl single-pressureet; 1s; outlet; economizer 2s: evaporatorsuperheated-steam outlet; circ. 2s: pumpevaporator plant, inlet; 2”s:circ. of whichevaporatorpump theinlet; outlet;scheme 2”s: is 3s: steam turbine inlet; 3sd: SD saturated steam outlet; 4s: steam turbine outlet. shownevaporator in Figure outlet; 3. 3s: steam turbine inlet; 3sd: SD saturated steam outlet; 4s: steam turbine outlet.

Figure 3.3. WHR single-pressure superheated-steam plant. AbbreviationsAbbreviations; ;the the same same of of Figure Figure 22 withwith additions; SH: superheater; 1g: superheater gas inlet; 3”s: superheater steam outlet. additions; SH: superheater; 1g: superheater gas inlet; 3”s: superheater steam outlet.

The basic layout of the superheated-steam system was the same as that adopted by the saturated- steam system, but with the addition of a superheater (SH in Figure 3) in the HRSG that takes saturated steam from the SD and feeds the steam turbine with superheated steam.

5. WHR Steam-Plant Simulation Models In order to define the design characteristics of the selected WHR steam plants and evaluate their performance, we used two computer codes that we developed in MATLAB®, as reported in [24]. The first code is able to calculate the dimensions, weight, and performance of the steam-plant components (HRSG economizer, evaporator, superheater, steam drum, steam turbine, hot well tank, etc.) in design working conditions, starting from available data of the DF engine (such as mass-flow rate and temperature of exhaust gas and scavenge air). The second code is used for the performance simulation of the WHR system (and relative components) in off-design working conditions. Starting from the first code, the WHR steam-plant design procedure was based on steady-state continuity and energy equations: ()=  Mi - Mo 0 (1) ()M h -M h = Q' - P  i i o o (2)

Energies 2020, 13, 985 6 of 21

5. WHR Steam-Plant Simulation Models In order to define the design characteristics of the selected WHR steam plants and evaluate their performance, we used two computer codes that we developed in MATLAB®, as reported in [24]. The first code is able to calculate the dimensions, weight, and performance of the steam-plant components (HRSG economizer, evaporator, superheater, steam drum, steam turbine, hot well tank, etc.) in design working conditions, starting from available data of the DF engine (such as mass-flow rate and temperature of exhaust gas and scavenge air). The second code is used for the performance simulation of the WHR system (and relative components) in off-design working conditions. Starting from the first code, the WHR steam-plant design procedure was based on steady-state continuity and energy equations: X (M Mo) = 0 (1) i − X (M h Moho) = Q0 P (2) i i − − where Mi and Mo are the fluid-mass-flow rates at the plant components’ inlet and outlet sections, respectively; h the fluid-specific enthalpy; and Q’ and P the heat and mechanical power exchanged with the outside, respectively. The heat exchangers of the considered HRSG were characterized by cross-counter flow configuration with horizontal finned tubes. The overall heat-exchange coefficient (ke) of the pipe wall for the various HRSG components was calculated as follows: 1 1 s Ae Ae 1 Ae = + Re + + Ri + (3) ke he k Aml Ai hi Ai

In this equation, he and hi are external and internal pipe-convective heat-transfer coefficients; Re and Ri are external and internal pipe thermal resistance; s and k are pipe-wall thickness and wall thermal conductivity, respectively; and Ae, Ai, and Aml are pipe wall external, internal, and logarithmic area, respectively. Convective heat-transfer coefficients he and hi were determined by using finned tube correlations [24]. For each HRSG heat exchanger, the total heat-exchange area (necessary for the pipe length evaluation) was calculated by means of heat-exchange balance equation:

Q0g = kgAe(∆T)g w (4) − which was relative to the gas-wall heat exchange, and:

Qs0 = ksAi(∆T)w s (5) − which was relative to the wall–steam heat exchange. In Equations (4) and (5), terms ∆Tg-w and ∆Tw-s, representing fluid–pipe-wall logarithmic temperature differences, are given, respectively, by:

To s Ti s (∆T) = − − − (6) w s Tw Ti s − ln − − Tw To s − − and: To s Ti s (∆T)g w = − − −  (7) − Ti g Tw ln − − To g Tw − − A similar procedure was used for the design of other heat exchangers of the steam plant: condenser (SCO), jacket water (JW), and scavenge air preheater (SC). The steam-turbine simulation model was based on typical steady-state nondimensional mass-flow maps [31]. The efficiency of the steam turbine was evaluated by a correlation that took into account steam quality, as suggested by a steam-turbine manufacturer (Fincantieri S.p.A.). The HRSG steam Energies 2020, 13, 985 7 of 21 drum (SD) and hot well tank (HWT) were modelled as simple fluid-mixing tanks. The engine back-pressure value due to the HRSG was determined as reported in [24]. As mentioned before, a second computer code was developed to determine WHR steam-plant performance at both design and off-design steady-state load conditions. Starting from the design data of the plant components (geometric and functional characteristics), this second simulation code determined the WHR plant performance by using a calculation procedure reversed in some way with respect to the design one, as explained in [24]. WHR steam-plant design and off-design simulation procedures were validated in other papers by the authors [24,29] by comparing simulation results with existing corresponding experiment data. The comparison between calculated parameters and reference data was generally good for WHR steam plants operating under both design conditions (differences generally no more than 1%) and off-design conditions (differences generally less than 4%). As an example of the validity of the presented simulation procedures, Table2 shows a comparison taken from [24] that shows differences in terms of percentage errors between simulation results and experiment data concerning a dual-pressure WHR plant operating in design conditions, applied to a specific propulsion system consisting of a MAN two-stroke DE (6S70ME-C8.2), selected as prime mover of a 158,000 DWT crude-oil tanker (Suezmax), belonging to the Premuda company.

Table 2. Accuracy of developed design procedure [23].

Design-Code Results % Error

HPTi: turbine inlet high pressure 0.01 LPTi: turbine inlet low pressure 0.00 MS HP: turbine inlet steam mass-flow rate 0.05 MS LP: turbine inlet steam mass-flow rate 0.52 ToutG: HRSG outlet gas temperature 0.54 PST: steam turbine power 0.75

6. WHR Steam-Plant Optimization The adopted optimization model, based on the calculation procedures described above, is similar to that described in [24]. In this study, the optimization procedure starts from the consideration that DF engines predominantly work powered by NG, and only occasionally use diesel fuel [26]. Therefore, performance optimization of the Dual Fuel engine–WHR Combined Plant (DF–WHR CP) had to consider the engine powered by NG. Once defined the WHR steam-plant layouts (see Figures2 and3), the operative engine NCR (488 rpm and 13,162.5 KW), and the SSS steam mass-flow rate, the Rankine cycle parameters (i.e., pressure, temperature, water/steam mass-flow rates), in the various WHR plant components, were thermodynamically optimized in order to maximize the DF–WHR CP efficiency (ηCP), defined as:

PDF E + PST PMFP PSCP ηCP = − − (8) M f LHV f where PDF E is the DF engine power; PST the steam turbine power; PMFP and PSCP the main feed and steam condenser pump powers; and Mf and LHVf the fuel mass-flow rate and the lower heating value, respectively. The WHR steam-plant optimization procedure starts from the definition of a series of fixed parameters, reported in Table3, referring to the DF engine operating at NCR conditions, fueled by NG, and assuming a vertical configuration for the HRSG. Energies 2020, 13, 985 8 of 21

Table 3. Fixed parameters of WHR system referring to DF engine operating at 75 % of MCR load.

Fixed WHR Plant Parameters NG Fuel Engine exhaust gas mass flow rate (kg/s) 26 Engine exhaust gas temperature (◦C) 352 Condenser pressure (bar) 0.08 Plan dimensions of the HRSG (m m) 1.85 2.10 × ×

Once the fixed parameters in Table3 were defined, the WHR steam plants were optimized by varying the HRSG optimization parameters reported in Table4, with the aim of maximizing the ηCP parameter (Equation (8)). The efficiency of the DF-WHR combined plant ηCP is in fact the objective function to be optimized in the optimization process concerning the HRSG parameters shown in Table4.

Table 4. Variation range and variation steps of HRSG optimization parameters.

HRSG Optimization Parameters Variation Range Variation Steps HRSG steam drum pressure (bar) 5–20 0.1 Pinch point temp. differences: ∆T pp (◦C) 5–20 1 Approach point temperature diff.: ∆T ap 5–35 1 (◦C) for superheated steam plant

The main results of the optimization procedure carried out for the WHR steam plants combined with the DF engine are shown in the following Tables5 and6 relative to the DF engine powered by NG or HFO, respectively. In both Tables the percentage differences reported in the fourth column were calculated as follows: x x ∆x/x% = SH SP − S SP 100[%] (9) xS SP where x represents the considered parameter and suffixes SH SP and S SP indicate the superheated- and saturated-steam plant, respectively.

Table 5. Optimization of WHR steam plants combined with DF engine powered by NG.

NG Fuel Optimized HRSG Parameters WHR Saturated WHR Superheated WHR Steam Steam Difference (%) HRSG steam drum pressure (bar) 7.6 8 5.26 Pinch point temp. difference: ∆T pp ( C) 5.1 5 1.96 ◦ − Approach point temp difference: ∆T ap (◦C) - 17 - Optimized Main HRSG Data HRSG height [m] 4.33 5.04 16.26 HRSG weight [t] 28.75 33.26 15.67 Economizer pipe number 7 6 14.29 − Evaporator pipe number 34 34 0.0 Superheater pipe number - 7 -

The obtained optimization results in case of the engine powered by NG are reported in the second and third columns of Table5 for saturated and superheated steam, respectively, while percentage differences are shown in the fourth column. The upper and lower part of the table refer to the thermodynamic and constructive parameters of the HRSG, respectively. Energies 2020, 13, 985 9 of 21

Table 6. Main HRSG data and WHR plant data in DF engine powered by Heavy Fuel Oil (HFO).

HFO Fuel HRSG Parameters WHR Saturated WHR Superheated WHR Steam Steam Difference (%) HRSG steam drum pressure (bar) 7.1 7.3 2.81 Pinch point temp. difference: ∆T pp (◦C) 7.2 7.9 9.72 Approach point temp difference: ∆T ap (◦C) - 7.3 - WHR Plants Data

ST inlet pressure (% of pSD) 70 70 0.0 HRSG gas outlet temperature (◦C) 165.3 169.3 2.38

Table6 shows the HRSG parameters in the case of the DF engine powered by HFO, running at NCR, obtained taking into account the temperature and mass-flow rates of the exhaust gases generated by this fuel (see Figure1). In this case, the geometrical data of the HRSG components were the same, determined by the optimization procedure relative to the engine powered by NG (see Table5). Similar to the case of Table5, the data reported in the second and third column of Table6 refer to saturated- and superheated-steam WHR plants, respectively. In order to satisfy the already mentioned constraint of steam pressure for ship steam services (at least 7 bar of HRSG SD pressure), it is necessary to partially close the Isenthalpic Lamination Valve (ILV in WHR plant schemes of Figures2 and3) to reduce pressure in the steam turbine inlet section to 70% of SD pressure, as reported in Table6. Table6 shows that, for the engine running at NCR, both WHR plants satisfied the constraint that gas temperature at the HRSG outlet should remain over the 160 ◦C limit, as required by the engine manufacturers when HFO is used as fuel due to the sulfur present in it, which causes acid corrosion at lower temperatures. The same temperature limit was also satisfied in the other considered engine loads.

7. WHR Steam-Plants Comparison A first comparison between saturated-steam (Figure2) and superheated-steam (Figure3) WHR plants is shown in the fourth column of Tables5 and6, referring to NG and HFO engine fuel, respectively. The comparison data reported below are relative to the engine running at NCR load conditions. Percentage differences relative to optimized HRSG parameters, reported in the fourth column of Table5 (NG fuel), showed small di fference (5.26%) for HRSG SD pressure, and even smaller difference ( 1.96%) for pinch point temperature difference (∆Tpp). − Differences of the same parameters were 2.81% and 9.72% in the case of HFO fuel (fourth column of Table6), respectively. Greater di fference was observed for main optimized HRSG data, shown in the fourth column of Table5, regarding HRSG height and weight (both about 16%). Figure4 shows details of saturated- and superheated-steam HRSG plants in terms of height (Figure4a) and weight (Figure4b) for HRSG components (economizer, evaporator, superheater, and shell). No considerable differences could be found between saturated- and superheated-steam plants with regard to the weight, height, and number of pipes (see Table5) of the economizers and evaporators. Figure4 shows that the greater height and weight of the superheated-steam plant was mainly due to the superheater. Figure4a shows the considerable shell contribution in overall HRSG height, while its contribution in total HRSG weight was small, as reported in Figure4b. Energies 2019, 12, x FOR PEER REVIEW 9 of 20

HFO Fuel WHR HRSG Parameters WHR Saturated WHR Superheated Difference Steam Steam (%) HRSG steam drum pressure (bar) 7.1 7.3 2.81 Pinch point temp. difference: ΔT pp 7.2 7.9 9.72 (°C) Approach point temp difference: ΔT ap - 7.3 - (°C) WHR Plants Data

ST inlet pressure (% of pSD) 70 70 0.0 HRSG gas outlet temperature (°C) 165.3 169.3 2.38

Table 6 shows that, for the engine running at NCR, both WHR plants satisfied the constraint that gas temperature at the HRSG outlet should remain over the 160 °C limit, as required by the engine manufacturers when HFO is used as fuel due to the sulfur present in it, which causes acid corrosion at lower temperatures. The same temperature limit was also satisfied in the other considered engine loads.

7. WHR Steam-Plants Comparison A first comparison between saturated-steam (Figure 2) and superheated-steam (Figure 3) WHR plants is shown in the fourth column of Tables 5 and 6, referring to NG and HFO engine fuel, respectively. The comparison data reported below are relative to the engine running at NCR load conditions. Percentage differences relative to optimized HRSG parameters, reported in the fourth column of Table 5 (NG fuel), showed small difference (5.26%) for HRSG SD pressure, and even smaller difference (–1.96%) for pinch point temperature difference (ΔT pp). Differences of the same parameters were 2.81% and 9.72% in the case of HFO fuel (fourth column of Table 6), respectively. Greater difference was observed for main optimized HRSG data, shown in the fourth column of Table 5, regarding HRSG height and weight (both about 16%). Figure 4 shows details of saturated- and superheated-steam HRSG plants in terms of height Energies(Figure2020 4a), 13 and, 985 weight (Figure 4b) for HRSG components (economizer, evaporator, superheater,10 ofand 21 shell).

50 45 SHELL ECO 40 EVA 35 SH 30 25 20 Weight [t] Weight Height [m] 15 10 5 0 12 Saturated Superheated (a) (b)

FigureFigure 4.4. ((a)a) HeightHeight andand ((b)b) weightweight ofof saturated-saturated- andand superheated-steamsuperheated-steam HRSGHRSG plantplant components.components. 7.1. WHR Steam Plants Performance in Design Working Conditions No considerable differences could be found between saturated- and superheated-steam plants withConsiderations regard to the similarweight, to height, those expressed and number in the of comment pipes (see of Figure Table4 can5) of be drawnthe economizers from the more and detailedevaporators. HRSG Figure data reported4 shows that in Table the 7greater. height and weight of the superheated-steam plant was Analysis of the most important data reported in Table7 led to the following considerations: Below are the formulas used to calculate the efficiency of the HRSG and its heat exchangers. With reference to Figures2 and3 and to the captions of Figures2 and3, these formulas are:

T1g T4g HRSG = − 100[%] (10) eff T T 1g − amb   T T 2g − 3g EVeff =   100 [%] (11) T T 2g − 2s   T T 3g − 4g Eeff =  100 [%] (12) T T 3g − 00SC   T T 1g − 2g SHeff =  100 [%] (13) T T 1g − 3sd From these formulas, we understand how EV efficiency was higher than that of E and the SH, as is easy to deduce from Figures 6 and 7. Despite less superheater efficiency (12.43% for NG and 6.14% for HFO, line 17 in Table7), the exchanged superheater thermal power (line 14) reduced the evaporators’ exchanged thermal powers by about 15% for NG and 10% for HFO (line 10), being practically equal the evaporator efficiency in the saturated- and superheated-steam plants (line 13). In the superheated-steam plant, when compared to the saturated one, the less exchanged evaporator thermal power (about 15% (NG) and 10% (HFO)) involved similar (about 15% and 10%, respectively) economizer water-mass flow-rate reduction (line 5 in Table7), able to maintain about the same steam drum pressure between the saturated and superheated plants, as reported in Table5 (NG) and Table6 (HFO). Similar less economizer efficiency (line 9 in Table7) in the superheated plant, compared to the saturated one, was observed for both NG and HFO engine fuels. Greater HRSG gas-side pressure loss (more than 16% for both considered engine fuels) was observed in the superheated plant compared to the saturated one (line 3 of Table7); that, however, did not substantially reduce engine e fficiency. Energies 2020, 13, 985 11 of 21

Table 7. Saturated- and superheated-steam HRSG parameters and percentage differences in engines powered by NG or HFO.

NG Fuel HFO Fuel HRSG Parameters WHR WHR WHR WHR WHR Sat. WHR Sat. Superh. Diff. (%) Superh. Diff. (%) Steam Steam Steam ∆x/x 100 Steam ∆x/x 100 1 Overall HRSG (%) 58.37 56.85 2.61 48.40 46.95 2.99 eff − − 2 HRSG gas outlet temperature (◦C) 160.3 165.2 3.08 165.3 169.3 2.38 3 HRSG gas side pressure loss (mbar) 30.01 34.94 16.46 43.03 50.27 16.52 4 Exchanged E thermal power (MW) 0.36 0.32 10.66 0.26 0.24 7.15 − − 5 E water-mass flow rate (kg/s) 2.27 1.91 15.85 2.44 2.18 10.65 − − 6 E water inlet temperature (◦C) 121.1 125.3 3.47 134.2 136.8 1.94 7 E water outlet temperature (◦C) 157.8 161.7 2.47 165.2 166.1 0.54 8 E gas inlet temperature (◦C) 174.1 177.8 2.16 173.4 176.8 2.00 9 E efficiency (%) 26.26 21.87 16.72 20.96 18.03 13.94 − − 10 EV exchanged thermal power (MW) 4.77 4.04 15.35 4.06 3.64 10.18 − − 11 EV steam drum temperature (◦C) 168.3 170.4 1.24 165.6 166.7 0.66 12 EV gas inlet temperature ( C) 350 328 6.29 297.0 287.6 3.17 ◦ − − 13 EV efficiency (%) 97.19 96.46 0.75 94.46 93.80 0.71 − − 14 SH exchanged thermal power (MW) - 0.62 - - 0.27 - 15 SH steam outlet temperature (◦C) - 332.9 - - 289.7 - 16 SH gas inlet temperature (◦C) - 350.0 - - 297.0 - 17 SH efficiency (%) - 12.43 - - 6.14 - Energies 2019, 12, x FOR PEER REVIEW 11 of 20 Energies 2019, 12, x FOR PEER REVIEW 11 of 20 byby about about 15% 15% for for NG NG and and 10% 10% for for HFO HFO (line (line 10), 10), bein being gpractically practically equal equal the the evaporator evaporator efficiency efficiency in in thethe saturated- saturated- and and superheated-steam superheated-steam plants plants (line (line 13). 13). InIn the the superheated-steam superheated-steam plant, plant, when when compared compared to to the the saturated saturated one, one, the the less less exchanged exchanged evaporatorevaporator thermal thermal power power (about (about 15% 15% (NG) (NG) and and 10% 10% (HFO)) (HFO)) involved involved similar similar (about (about 15% 15% and and 10%, 10%, respectively)respectively) economizer economizer water-mass water-mass flow-rate flow-rate reduct reductionion (line (line 5 5in in Table Table 7), 7), able able to to maintain maintain about about thethe same same steam steam drum drum pressure pressure between between the the saturate saturated dand and superheated superheated plants, plants, as as reported reported in in Table Table 5 5 (NG)(NG) and and Table Table 6 6(HFO). (HFO). Similar Similar less less economizer economizer efficiency efficiency (line (line 9 9in in Table Table 7) 7) in in the the superheated superheated plant,plant, compared compared to to the the saturated saturated one, one, was was observ observeded for for both both NG NG and and HFO HFO engine engine fuels. fuels. Greater Greater HRSGHRSG gas-side gas-side pressure pressure loss loss (more (more than than 16% 16% for for both both considered considered engine engine fuels) fuels) was was observed observed in in the the superheatedEnergiessuperheated2020, 13 ,plant 985plant compared compared to to the the saturated saturated on one e(line (line 3 3of of Table Table 7); 7); that, that, however, however, did did12 ofnot not 21 substantiallysubstantially reduce reduce engine engine efficiency. efficiency. The saturated- and superheated-steam plants were characterized by very similar overall The saturated-saturated- and and superheated-steam superheated-steam plants plants were were characterized characterized by very by similar very overallsimilar e ffioverallciency, efficiency, as reported in line 1 (HRSGeff) of Table 7, with an advantage of near 10% in the case of NG efficiency,as reported as inreported line 1 (HRSGin line 1 )(HRSG of Tableeff) 7of, withTable an 7, with advantage an advantage of near of 10% near in 10% the in case the of case NG of fuel NG fuel compared to HFO fuel.e Theff same considerations can be deduced from Figure 5, where available fuelcompared compared to HFO to HFO fuel. fuel. The sameThe same considerations considerations can be can deduced be deduced from Figurefrom Figure5, where 5, availablewhere available HRSG HRSG gas thermal power is also reported in the third column. HRSGgas thermal gas thermal power power is also reportedis also reported in the thirdin the column. third column.

10 10 SH SH EVA EVA 8 ECO 8 ECO

6 6

4 4

2 Thermal power [MW]powerThermal 2 [MW] power Thermal Thermal power [MW]powerThermal Thermal power [MW] power Thermal

0 0 123 123 1-Saturated 2-Superheated 3-AV Thermal power 1-Saturated 2-Superheated 3-AV Thermal power (a()a ) (b()b )

FigureFigure 5. 5. Engine Engine exhaust exhaust gas gas thermal thermal power power (3) (3) and and ther thermal thermalmal power power recovered recovered by by (1) (1) saturated- saturated- and and (2)(2) superheated-steam superheated-steam plant plant plant heat heat heat exchangers: exchangers: exchangers: E, E, E, EV, EV, EV, and and and SH, SH, SH, with with with an an an engine engine engine powered powered powered by by by ( (a a())a NG NG) NG or or or (b()b HFO. ) HFO.

FigureFigure 66 shows6shows shows thethe the temperaturetemperature temperature profilesprofiles profiles ofof of gasga gas andsand and waterwater/steam water/steam/steam flows flows flows along along the the HRSG HRSG versus versus wastewaste heat heat recovery recovery for for the the saturated-steam saturated-steam plant plant with with an an engine engine fed fed by by (a) (a) NG NG or or (b) (b) HFO. HFO. The The same same isis shown shown in in Figure Figure 77 for7for for thethe the superheated-steamsuperheated-steam superheated-steam plant.plant. plant.

400 400 HRSG water-steam HRSG water-steam 350 HRSG exhaust gas 350 HRSG exhaust gas scavenger water 1g = 2g 300 scavenger water 1g = 2g 300 scavenger air scavenger air 250 250 SC SC ai 200 ai 3g 200 3g 3s 4g 3s 150 4g 150 00sc 2s 00sc 1s2s 100 1s 100 0sc 0sc SC SC ao 50 ao 50 SC EEV 0SC EEV HRSG gas and steam temp. [°C] 0 HRSG gas and steam temp. [°C] 0 102030405060708090100 0 102030405060708090100 Waste heat recovered [%] Waste heat recovered [%] (a()a ) (b()b )

FigureFigure 6. 6. TemperatureTemperature Temperature profiles profiles profiles along along HRSG HRSG for for saturated- saturated-steam saturated-steamsteam plant plant with with with an an an engine engine engine powered powered powered by by by ( (aa ())a ) NGNG or or ( b ()b HFO. ) HFO. Further information on the engine, steam plants, and the DF engine–WHR combined system can be found in Table8. Line 1 of Table8 shows that, with the engine powered by NG, HRSG back

pressure for both types of steam plant increased the engine-specific fuel consumption by only about 1 g, and by slightly more with the engine powered by HFO. With both types of fuel, engine efficiency remained substantially the same for the saturated- and the superheated-steam plant (line 2 in Table8). Line 3 of Table8 shows that Rankine cycle e fficiency was very similar in the cases of the engine powered by NG or HFO; a greater cycle efficiency of 6.91% was observed in the case of NG fuel for the superheated-steam plant compared to the saturated-steam one, and this difference dropped to 4.41% when using HFO. Energies 2020, 13, 985 13 of 21

Table 8. Engine, steam-plant components, and DF engine–WHR combined plants: main data and percentage differences in engine powered by NG or HFO.

Engine, Steam-Plant Components, and NG Fuel HFO Fuel Combined Plant Data WHR WHR WHR Sat. WHR Superh. WHR Sat. WHR Diff. (%) Diff. (%) Superh. Steam Steam Steam ∆x/x 100 ∆x/x 100 Steam

1 HRSG p loss. Esfc incr. (g/kWh) 0.92 1.07 16.46 1.31 1.52 16.50 2 η (%) 47.28 47.23 0.10 45.67 45.61 0.12 E − − 3 ηRankine (%) 26.16 27.96 6.91 25.80 26.91 4.41 4 ηCP (%) 50.88 51 0.25 47.93 47.95 0.05 5 NCR engine power (kW) 13162.5 13162.5 0.0 13162.5 13162.5 0.0 6 ST power (kW) 1014.0 1062.5 4.81 657.2 679.6 3.41 7 ST inlet pressure (% of pSD) 100 100 0.0 70 70 0.0 8 ST ∆Hid (kJ/kg) 678.2 825.7 21.76 494.1 570.4 15.43 9 ST mass flow rate (kg/s) 2.10 1.74 17.38 1.87 1.61 13.90 − − 10 ST efficiency (%) 71.0 74.0 4.18 71.2 74.0 3.89 11 ST outlet steam quality 0.87 0.97 12.00 0.87 0.95 9.40 12 condenser thermal power (kW) 4390 4074.6 7.18 3918.7 3591.1 5.81 − − 13 SCP pump power (kW) 17.63 16.32 7.46 11.63 10.98 5.57 − − 14 MFP pump power (kW) 4.51 4.99 10.8 3.58 4.16 16.24 15 SSS mass flow rate (kg/s) 0.17 0.17 0.0 0.57 0.57 0.0 Energies 2020, 13, 985 14 of 21 Energies 2019, 12, x FOR PEER REVIEW 12 of 20

(a) (b)

FigureFigure 7. 7.Temperature Temperature profiles profiles along along HRSG HRSG for for superheated-steam superheated-steam plant plant with with an an engine engine powered.by powered.by (a()a NG) NG or or (b ()b HFO.) HFO.

LineFurther 4 of information Table8 shows on the about engine, the samesteam DFplants engine–WHR, and the DF combinedengine–WHR plant combined e fficiency system (ηCP can), determinedbe found in by Table Equation 8. Line (8), in1 of the Table case of8 shows NG fuel. that, A similarwith the result engine was powered observed by in theNG, case HRSG of HFO back fuel,pressure although for both with types a diff erenceof steam of plant about increased 3% in favor the of engine-specific NG. The slight fuel ST consumption power difference by only between about the1 gram, saturated- and andby slightly superheated-steam more with the plants engine that po occurredwered by with HFO. both With tested both fuel types types, of as fuel, shown engine in lineefficiency 6 of Table remained8, was duesubstantially to a greater the STsame enthalpy for the dropsaturated- in the and case the of superheated-steam the superheated-steam plant plant (line 2 (linein Table 8); that, 8). however,Line 3 of Table was approximately 8 shows that Rankine compensated cycle efficiency by greater was ST very mass-flow similar rate in the (line cases 9) in of the the caseengine of the powered saturated-steam by NG or plant. HFO; a greater cycle efficiency of 6.91% was observed in the case of NG fuelTable for 8the shows superheated-steam that ST power plant with thecompared engine to powered the saturated-steam by HFO was aboutone, and 35% this lower difference than thatdropped with theto 4.41% engine when powered using byHFO. NG. This is mainly due to less ST enthalpy drop (line 8) in the case of HFO fuel, motivated by the need to partially close the ILV valve (Table6) to maintain SD pressureTable at 78. barEngine, at least, steam-plant as required components, by ship and steam DF engine–WHR services. In combined the case plants: of the main engine data powered and by HFO,percentage greater differences steam-mass-flow in engine rate powered was alsoby NG required or HFO. by ship steam services (line 15 of Table8) compared to the case of the engine powered by NG, contributing to a decrease of ST mass-flow rate, NG Fuel HFO Fuel and therefore, of its power. ST efficiencies were similar (line 10 in Table8) with a little advantage for WHR the superheated-steamEngine, Steam-Plant plant due to greater steam quality at the ST outlet (line 11). The powers of the WHR WHR Diff. WHR WHR SCP andComponents, MFP pumps, and much Combined lower than the ST power (about 0.8% andWHR 0.2%, Sat. respectively), as shown in Sat. Superh. (%) Superh. Diff. (%) lines 6, 13, andPlant 14 of TableData 8, had practically no influence on e fficiencySteamη of the DF engine–WHR Steam Steam Δx/x CP Steam Δx/x 100 combined plant (see Equation (8)). 100 7.2. 1 WHRHRSG Steam-Plants p loss. Esfc incr. Performance (g/kWh) in Off0.92-Design Working1.07 Conditions16.46 1.31 1.52 16.50 2 ηE (%) 47.28 47.23 –0.10 45.67 45.61 –0.12 In this section, we compare the DF engine–WHR combined plants in the cases of an engine powered 3 ηRankine (%) 26.16 27.96 6.91 25.80 26.91 4.41 by NG or HFO, as well as for different engine loads: 50%, 75% (adopted for WHR plant-component 4 ηCP (%) 50.88 51 0.25 47.93 47.95 0.05 design optimization), 85%, and 100% of the engine MCR. Figure8 and the following show the di fferent 5 NCR engine power (kW) 13162.5 13162.5 0.0 13162.5 13162.5 0.0 parameters that define the DF engine and the components of the WHR steam plants, highlighting the 6 ST power (kW) 1014.0 1062.5 4.81 657.2 679.6 3.41 percentage differences between the parameters of the superheated- and the saturated-steam plant. 7 ST inlet pressure (% of pSD) 100 100 0.0 70 70 0.0 These percentage differences, calculated with Equation (9), are displayed for the different engine loads 8 ST ΔHid (kJ/kg) 678.2 825.7 21.76 494.1 570.4 15.43 mentioned. In the figures, the horizontal lines signed with 0 represent zero-percent difference values. 9 ST mass flow rate (kg/s) 2.10 1.74 –17.38 1.87 1.61 –13.90 As shown in Figure8a,b, pertinent to NG and HFO fuel, respectively, the delivered engine power 10 ST efficiency (%) 71.0 74.0 4.18 71.2 74.0 3.89 (DF Engine ) was always maintained constant regardless of the used WHR plant. With regard to the 11 STbr outlet steam quality 0.87 0.97 12.00 0.87 0.95 9.40 Steam Condensed Pump and the Main Feed Pump (SCP and MFP in the schemes of Figures2 and3), 12 condenser thermal power (kW) 4390 4074.6 –7.18 3918.7 3591.1 –5.81 the power differences between the two compared WHR plants and types of fuel generally increased as 13 SCP pump power (kW) 17.63 16.32 –7.46 11.63 10.98 –5.57 engine load decreased, especially in the case of the MFP pump (Figure8a). However, these di fferences 14 MFP pump power (kW) 4.51 4.99 10.8 3.58 4.16 16.24 had negligible influence on the engine–WHR combined plant efficiency (Equation (8)), as their power 15 SSS mass flow rate (kg/s) 0.17 0.17 0.0 0.57 0.57 0.0 was much lower than ST power. Line 4 of Table 8 shows about the same DF engine–WHR combined plant efficiency (ηCP), determined by Equation (8), in the case of NG fuel. A similar result was observed in the case of HFO fuel, although with a difference of about 3% in favor of NG. The slight ST power difference between

Energies 2019, 12, x FOR PEER REVIEW 13 of 20

the saturated- and superheated-steam plants that occurred with both tested fuel types, as shown in line 6 of Table 8, was due to a greater ST enthalpy drop in the case of the superheated-steam plant (line 8); that, however, was approximately compensated by greater ST mass-flow rate (line 9) in the case of the saturated-steam plant. Table 8 shows that ST power with the engine powered by HFO was about 35% lower than that with the engine powered by NG. This is mainly due to less ST enthalpy drop (line 8) in the case of HFO fuel, motivated by the need to partially close the ILV valve (Table 6) to maintain SD pressure at 7 bar at least, as required by ship steam services. In the case of the engine powered by HFO, greater steam-mass-flow rate was also required by ship steam services (line 15 of Table 8) compared to the case of the engine powered by NG, contributing to a decrease of ST mass-flow rate, and therefore, of its power. ST efficiencies were similar (line 10 in Table 8) with a little advantage for the superheated- steam plant due to greater steam quality at the ST outlet (line 11). The powers of the SCP and MFP pumps, much lower than the ST power (about 0.8% and 0.2%, respectively), as shown in lines 6, 13, and 14 of Table 8, had practically no influence on efficiency ηCP of the DF engine–WHR combined plant (see Equation 8).

7.2. WHR Steam-Plants Performance in Off-Design Working Conditions In this section, we compare the DF engine–WHR combined plants in the cases of an engine powered by NG or HFO, as well as for different engine loads: 50%, 75% (adopted for WHR plant- component design optimization), 85%, and 100% of the engine MCR. Figure 8 and the following show the different parameters that define the DF engine and the components of the WHR steam plants, highlighting the percentage differences between the parameters of the superheated- and the saturated-steam plant. These percentage differences, calculated with Equation (9), are displayed for the different engine loads mentioned. In the figures, the horizontal lines signed with 0 represent zero- percent difference values. As shown in Figure 8a and 8b, pertinent to NG and HFO fuel, respectively, the delivered engine power (DF Enginebr) was always maintained constant regardless of the used WHR plant. With regard to the Steam Condensed Pump and the Main Feed Pump (SCP and MFP in the schemes of Figures 2 and 3), the power differences between the two compared WHR plants and types of fuel generally increased as engine load decreased, especially in the case of the MFP pump (Figure 8a). However, these differences had negligible influence on the engine–WHR combined plant efficiency (Equation Energies 2020, 13, 985 15 of 21 (8)), as their power was much lower than ST power.

(a) (b)

FigureFigure 8. 8.Engine Engine and and WHR WHR plant plant components: components: power power versus versus engine engine load load with with an an engine engine powered powered by by (a()a) NG NG or or (b (b) HFO.) HFO.

TheThe same same figure figure shows shows a a little little advantage advantage (always (always less less than than 10%) 10%) of of real real delivered delivered ST ST power power (ST (STr)r) ofof the the WHR WHR superheated-steam superheated-steam plant, plant, for for all all considered considered engine engine powers powers and and fuel fuel types. types. This This small small powerEnergiespower di2019 differencefference, 12, x FOR could could PEER be beREVIEW explained explained by by individual individual analysis analysis of of ST ST parameters, parameters, reported reported in in Figure Figure149 .of 9. 20

180 180 ST mass flow ST mass flow 150 ST isentropic DH 150 ST isentropic DH ST inlet temp. 120 120 ST inlet temp.

90 90 60 60 30 30 0 0 -30 Steam turbine parameters [%]

-30 Steam parameters turbine [%] 50 60 70 80 90 100 -60 Engine power [%] MCR 50 60 70 80 90 100 Engine power [%] MCR (a) (b)

FigureFigure 9.9. Steam-turbineSteam-turbine parametersparameters versus versus engine engine load load with with an an engine engine powered powered by by ( a(a)) NG NG or or ( (bb)) HFO. HFO.

Indeed,Indeed, from from Figure Figure9a, 9a, pertinent pertinent to NGto NG fuel, fuel, a greater a greater ST isentropic-enthalpy ST isentropic-enthalpy percentage percentage drop (DH) drop can(DH) be can observed be observed in the superheated in the superheated cycle, as opposedcycle, as toopposed an approximately to an approximately equal smaller equal percentage smaller masspercentage flow. This mass substantial flow. This equivalence substantial betweenequivalenc thee STbetween isentropic-enthalpy the ST isentropic-enthalpy percentage drop percentage increase anddrop the increase mass flow and percentagethe mass flow decrease percentage is due todecrea superheaterse is due presence to superheater in the HRSG, presence that in reduces the HRSG, the gasthat temperature reduces the enteringgas temperature the evaporator; entering it the follows evaporator; that to it obtain follows the that desired to obtain steam the drum desired pressure steam (verydrum similar pressure for (very saturated similar and for superheated saturated and steam supe plants),rheated the steam processed plants), water-steam the processed flow water-steam rate must beflow reduced, rate must compared be reduced, to the casecompared of the to saturated the case steam of the plant, saturated in which steam the plant, gas temperature in which the at thegas evaporatortemperature inlet at the is the evaporator same as atinlet the is outlet the same of the as engine at the turbocharger.outlet of the engine turbocharger. ThisThis STST parameterparameter compensationcompensation waswas lessless evidentevident inin thethe casecase ofof HFOHFO fuelfuel (Figure(Figure9 b),9b), as as is is indicatedindicated by by the the ST ST power power variation variation in in Figure Figure8 b.8b. FigureFigures 10 10aa,b showsand b shows a comparison a comparison of the saturated- of the saturated- and superheated-steam and superheated-steam WHR plants WHR in plants terms ofin HRSGterms efficiencyof HRSG as aefficiency whole and as of itsa components,whole and particularlyof its components, the evaporator particularly and economizer, the evaporator also including and theeconomizer, temperature also of theincluding exhaust the gas attemperature the HRSG outlet of the (T 4gexhaustin plant gas schemes at the of HRSG Figures outlet2 and 3(T). 4g in plant schemesAs can of beFigures seen from2 and the 3). figure, overall HRSG efficiency in the case of the superheated-steam plant was lower than that obtained in the case of the saturated-steam plant for all considered engine loads,

with the20 exception of the 50% engine load in the case of HFO20 engine fuel (Figure 10b). This isdue to Overall HRSG

[%] Overall HRSG very similar evaporator efficiency between the two compared [%] steam plants and fuel types, and greater 4g Evaporator 4g Evaporator Economizer 10 10 Economizer HRSG T 4g HRSG T 4g 0 0

-10 -10

HRSG components eff. & T eff. HRSG components -20 50 60 70 80 90 100 HRSG components eff. & T -20 Engine power [%] MCR 50 60 70 80 90 100 Engine power [%] MCR (a) (b)

Figure 10. HRSG and components efficiency comparison versus engine load with an engine powered by (a) NG or (b) HFO.

As can be seen from the figure, overall HRSG efficiency in the case of the superheated-steam plant was lower than that obtained in the case of the saturated-steam plant for all considered engine loads, with the exception of the 50% engine load in the case of HFO engine fuel (Figure 10b). This is due to very similar evaporator efficiency between the two compared steam plants and fuel types, and greater difference (about 18% with NG fuel, less in the HFO case) regarding economizer efficiency in favor of the saturated-steam plant. This last effect on overall HRSG efficiency was mitigated by the much lower exchange of thermal power of the economizer compared to that of the evaporator. The HRSG exhaust gas temperature

Energies 2019, 12, x FOR PEER REVIEW 14 of 20

180 180 ST mass flow ST mass flow 150 ST isentropic DH 150 ST isentropic DH ST inlet temp. 120 120 ST inlet temp.

90 90 60 60 30 30 0 0 -30 Steam turbine parameters [%]

-30 Steam parameters turbine [%] 50 60 70 80 90 100 -60 Engine power [%] MCR 50 60 70 80 90 100 Engine power [%] MCR (a) (b)

Figure 9. Steam-turbine parameters versus engine load with an engine powered by (a) NG or (b) HFO.

Indeed, from Figure 9a, pertinent to NG fuel, a greater ST isentropic-enthalpy percentage drop (DH) can be observed in the superheated cycle, as opposed to an approximately equal smaller percentage mass flow. This substantial equivalence between the ST isentropic-enthalpy percentage drop increase and the mass flow percentage decrease is due to superheater presence in the HRSG, that reduces the gas temperature entering the evaporator; it follows that to obtain the desired steam drum pressure (very similar for saturated and superheated steam plants), the processed water-steam flow rate must be reduced, compared to the case of the saturated steam plant, in which the gas temperature at the evaporator inlet is the same as at the outlet of the engine turbocharger. This ST parameter compensation was less evident in the case of HFO fuel (Figure 9b), as is indicated by the ST power variation in Figure 8b. Energies 2020, 13, 985 16 of 21 Figures 10a and b shows a comparison of the saturated- and superheated-steam WHR plants in terms of HRSG efficiency as a whole and of its components, particularly the evaporator and dieconomizer,fference (about also 18% including with NG the fuel, temperature less in the HFOof the case) exhaust regarding gas at economizer the HRSG e ffioutletciency (T in4g favorin plant of theschemes saturated-steam of Figures plant.2 and 3).

20 20 Overall HRSG

[%] Overall HRSG [%]

4g Evaporator 4g Evaporator Economizer 10 10 Economizer HRSG T 4g HRSG T 4g 0 0

-10 -10

HRSG components eff. & T eff. HRSG components -20 50 60 70 80 90 100 HRSG components eff. & T -20 Engine power [%] MCR 50 60 70 80 90 100 Engine power [%] MCR (a) (b)

FigureFigure 10. 10.HRSG HRSG and and components components e ffiefficiencyciency comparison comparison versus versus engine engine load load with with an an engine engine powered powered byby (a ()a NG) NG or or (b ()b HFO.) HFO.

EnergiesThisAs 2019 can last, 12 ebe,ff xect FORseen on PEER from overall REVIEW the HRSG figure, effi overallciency wasHRSG mitigated efficiency by in the the much case lower of the exchange superheated-steam of thermal15 of 20 powerplant was of the lower economizer than that compared obtained in to the that case of the of the evaporator. saturated-steam The HRSG plant exhaust for all considered gas temperature engine difference (HRSG T4g in Figure 9a) was always positive (greater value in superheated-steam WHR diloads,fference with (HRSG the exception T4g in Figure of the9a) 50% was engine always load positive in the (greatercase of HFO value engine in superheated-steam fuel (Figure 10b). WHR This is plant) and increased slightly as engine load decreased. plant)due to and very increased similar evaporator slightly as engineefficiency load between decreased. the two compared steam plants and fuel types, and In the case of HFO engine fuel (Figure 10b), this parameter difference was slightly negative only greaterIn the difference case of HFO (about engine 18% with fuel (FigureNG fuel, 10 lessb), this in the parameter HFO case) diff regardingerence was economizer slightly negative efficiency only in at 50% engine load. This was probably due to the ILV valve being partially closed for the reasons atfavor 50% of engine the saturated-steam load. This was plant. probably due to the ILV valve being partially closed for the reasons already mentioned. alreadyThis mentioned. last effect on overall HRSG efficiency was mitigated by the much lower exchange of thermal Data reported in Figure 11a and 11b regarding the efficiency comparison of engine, combined powerData of reportedthe economizer in Figure compared 11a,b regarding to that of the the effi evaporator.ciency comparison The HRSG of engine,exhaust combined gas temperature WHR plantsWHR andplants their and components their components showed showed less effi ciencyless effi ofciency the saturated of the saturated Rankine Rankine cycle compared cycle compared to the superheatedto the superheated one. This one. diff erenceThis difference increased increased as engine loadas engine decreased load in decreased the case of in NG the fuel case (Figure of NG 11 fuela), while(Figure it increased 11a), while with it increased an engine with load inan the engine case load of HFO in the fuel, case asshown of HFO in fuel, Figure as 11shownb, probably in Figure due 11b, to theprobably partial closingdue to the of the partial ILV valveclosing when of the engine ILV loadvalve decreased. when engine Steam-turbine load decreased. efficiency Steam-turbine was also aboutefficiency 4% greater was also in the about superheated-steam 4% greater in the plant superh for alleated-steam engine loads plant in the for case all engine of NG fuelloads (Figure in the 11 casea); whenof NG engine fuel (Figure fuel was 11a); HFO, when it remained engine fuel constant was HFO, at 4% it remained for high engine constant loads, at 4% and for increased high engine from loads, 4% toand 6% increased only for a from 50% MCR4% to engine6% only load. for a 50% MCR engine load.

(a) (b)

FigureFigure 11. 11.WHR WHR plant plant and and component-e component-efficiencyfficiency comparison comparison versus versus engine engine load load with with an an engine engine using using (a()a NG) NG or or (b ()b HFO.) HFO.

DespiteDespite this, this, the the di differencefference of of the the combined combined plant plant e ffiefficienciesciencies was was always always very very small small (with (with an an advantageadvantage for for the the superheated-steam superheated-steam plant plant mainly mainly in in the the case case of of NG NG fuel). fuel). This This was was due due to to greater greater engine power value compared to steam-turbine power, as reported in lines 5 and 6 of Table 8 for the NCR engine load. In fact, from Table 8 it can be estimated that the ST power is less than 8% of the engine power in case of NG fuel, and about 5% of the engine power in case of HFO.

8. Exergetic Analysis In the preceding sections, we made a comparison between the two WHR in light of the first principle of thermodynamics. More complete analysis should also consider the second principle, taking into account the irreversibility of the real processes, accompanied by entropy production corresponding to energy loss. In other words, the energy-recovery process should be evaluated not only from a quantitative but also from a qualitative point of view. To this aim, a useful approach is represented by exergetic analysis. The exergy of a system at a certain thermodynamic state may be defined as “the maximum amount of work that can be obtained when the system moves from that particular state to a state of equilibrium with the surroundings” [32]. For a thermomechanical process with stationary flow, specific exergy Ex can be calculated as follows:

Ex = (h – h0) – T0 (s – s0) (14)

where h and s are the specific enthalpy and entropy, and subscript 0 refers to ambient conditions. In our study, it was of particular interest to evaluate, for the considered process, the exergy efficiency, commonly defined as:

Energies 2020, 13, 985 17 of 21 engine power value compared to steam-turbine power, as reported in lines 5 and 6 of Table8 for the NCR engine load. In fact, from Table8 it can be estimated that the ST power is less than 8% of the engine power in case of NG fuel, and about 5% of the engine power in case of HFO.

8. Exergetic Analysis In the preceding sections, we made a comparison between the two WHR in light of the first principle of thermodynamics. More complete analysis should also consider the second principle, taking into account the irreversibility of the real processes, accompanied by entropy production corresponding to energy loss. In other words, the energy-recovery process should be evaluated not only from a quantitative but also from a qualitative point of view. To this aim, a useful approach is represented by exergetic analysis. The exergy of a system at a certain thermodynamic state may be defined as “the maximum amount of work that can be obtained when the system moves from that particular state to a state of equilibrium with the surroundings” [32]. For a thermomechanical process with stationary flow, specific exergy Ex can be calculated as follows:

Ex = (h h ) T (s s ) (14) − 0 − 0 − 0 where h and s are the specific enthalpy and entropy, and subscript 0 refers to ambient conditions. In our study, it was of particular interest to evaluate, for the considered process, the exergy efficiency, commonly defined as:

Use f ul Exergy f rom the process Wreal ηex = = (15) Total Exergy to the process He f where Wreal is the real work obtained by the process and Hef is the exergy heat amount delivered to the process by the fuel. It is possible to rewrite the previous equation on the basis of the first principle of thermodynamics:

Wreal Wreal/Qin ηex =   =   (16) Qin 1 T0 1 T0 − T − T

In the last term of the preceding equation, the numerator represents energy efficiency ηen (ratio between real work and incoming heat Qin), while the denominator is Carnot efficiency ηc. Thus, it is possible to write [32,33]: ηen ηex = (17) ηc The above-described formulation was applied to evaluate the considered WHR systems from an exergetic point of view as well. The involved thermodynamic processes are characterized by the following main input data, differing according to the selected type of WHR plant, the type of fuel used, and the engine-load conditions: engine exhaust gas temperature (T in Equation (16)); cooling-water temperature at condenser inlet (depending on the ambient conditions); and Rankine cycle efficiency (ηen in Equation (17)). The most important results of the developed analysis are shown in the following figures. Using HFO as engine fuel, Figure 12a shows that exergetic efficiency tendentially increased with decreasing engine power for both the saturated- and the superheated-steam plant. The variation of Rankine cycle efficiency was different with decreasing engine power. As shown in Figure 12b, a slight decrease of efficiency with decreasing engine power from 100% to 75% was followed by an increase when engine load decreased from 75% to 50%. Energies 2019, 12, x FOR PEER REVIEW 16 of 20

real ex = = (15)

where Wreal is the real work obtained by the process and Hef is the exergy heat amount delivered to the process by the fuel. It is possible to rewrite the previous equation on the basis of the first principle of thermodynamics:

real real/in = = ex 0 0 (16) (1 − ) (1 − ) in

In the last term of the preceding equation, the numerator represents energy efficiency ηen (ratio between real work and incoming heat Qin), while the denominator is Carnot efficiency ηc. Thus, it is possible to write [32,33]: en ex = (17) c The above-described formulation was applied to evaluate the considered WHR systems from an exergetic point of view as well. The involved thermodynamic processes are characterized by the following main input data, differing according to the selected type of WHR plant, the type of fuel used, and the engine-load conditions: engine exhaust gas temperature (T in Equation (16)); cooling-water temperature at condenser inlet (depending on the ambient conditions); and Rankine cycle efficiency (ηen in Equation (17)). The most important results of the developed analysis are shown in the following figures. Using HFO as engine fuel, Figure 12a shows that exergetic efficiency tendentially increased with decreasing engine power for both the saturated- and the superheated-steam plant. The variation of Rankine cycle efficiency was different with decreasing engine power. As shown in Figure 12b, a slight

Energiesdecrease2020 ,of13 ,efficiency 985 with decreasing engine power from 100% to 75% was followed by an increase18 of 21 when engine load decreased from 75% to 50%.

55 29 28 52.5 27 50 26

47.5 25 WHR sat. HFO WHR sat. HFO 24 WHR sat. NG WHR sat. NG 45 WHR superh. HFO 23 WHR superh. HFO WHR superh. NG WHR superh. NG 42.5 22 50 60 70 80 90 100 50 60 70 80 90 100 Engine power [%] MCR Engine power [%] MCR (a) (b)

FigureFigure 12. 12.WHR WHR steam-plant steam-plant ( a()a) exergetic exergetic and and ( b(b)) energetic energetic e ffiefficiencyciency variations variations versus versus engine engine power. power.

TheThe tendential tendential increase increase of of exergetic exergetic effi efficiencyciency with with decreasing decreasing engine engine load load from from 100% 100% to 75% to can75% becan explained be explained by considering by considering that thethat e ffitheciencies efficienci ofes both of Rankineboth Rankine and Carnotand Carnot cycle cycle decreased decreased at the at samethe same time (seetime Figure(see Figure1a). 1a). TheThe situation situation that that occurred occurred when when using using NG NG as as engine engine fuel fuel was was quite quite di different.fferent. InIn thisthis case,case, RankineRankine cycle cycle e ffiefficiency,ciency, as as shown shown in in Figure Figure 12 12b,b, remained remained almost almost constant constant when when varying varying engine engine load,load, while while Carnot Carnot e ffiefficiencyciency tended tended to to increase increase with with the the decrease decrease of of engine engine load load (see (see Figure Figure1a). 1a). This This explains the reduction of exergetic efficiency with the decrease of engine load, more marked in the case of the saturated-steam plant, when using NG as engine fuel.

9. Conclusions Using available simulation models, it was possible to carry out a detailed comparison between the considered saturated- and superheated-steam WHR plants, applied to the waste-heat recovery of the same DF marine engine. The comparison evidenced substantial equivalence between them with regards to the efficiency of the DF engine–WHR combined plants in all considered engine powers for the engine fueled by NG or HFO. A 3% higher efficiency is observed in the case of the engine powered by NG at normal continuous rating. The reasons of this substantial overall performance equivalence between single-pressure saturated- and superheated-steam WHR plants, with an advantage for the NG engine fuel, were analyzed in this paper in detail and were about the same for the two considered engine-fuel types. These reasons can be summarized in the substantial equivalence of the HRSG efficiency, with an advantage of about 10% in favor of the NG-fueled engine running at NCR, and in minor differences regarding the power of the steam turbine (never higher than 9%), always related to the same engine-fuel type. With regard to HRSG height and weight, the saturated-steam plant was characterized by a reduction of about 16% compared to the superheated-steam plant. The lower steam quality at the turbine outlet for the saturated-steam plant, compared to the superheated one, about equal for both engine fuels, was a disadvantage for the saturated-steam plant, because of the lower turbine efficiency and the greater damage of the turbine blades, both due to higher steam humidity in the saturated-steam cycles. To reduce erosion of turbine blades, determined by steam humidity, turbine manufacturers usually use more resistant blade materials to keep ST maintenance-time intervals about equal between saturated- and superheated-steam cycles. With regard to the overall costs of the considered WHR steam plants based on a saturated- or superheated-steam cycle, it can be said as a first approximation that they were approximately equal, the greater cost of the superheater in the HRSG in some way being compensated by the greater cost of the steam turbine operating with the saturated-steam cycle, because of the better material used for the Energies 2020, 13, 985 19 of 21 turbine blades. However, these economic aspects would require more in-depth analysis and could be developed in the near future. The authors believe that the research carried out and the obtained results may represent a contribution in the direction of increasing the efficiency of marine propulsion systems based on the use of dual-fuel (diesel fuel and natural gas) engines, characterized by a series of technical and environmental advantages. They also believe that the availability of suitable simulation procedures could be applied to optimize and compare the different solutions currently used or that are under development for the propulsion of commercial LNG carriers.

Author Contributions: The authors are part of a research group where two departments of the University of Genoa and one department of the University of Naples have collaborated on topics of common interest. Although the tasks entrusted to the individual participants have been appropriately differentiated (bibliographic research, development and validation of computational codes, comparison of results, writing of the article, subsequent revisions of it, etc.), it can be stated the substantial equivalence of the contribution that each author has provided, in quantitative and qualitative terms, for the preparation of this article. All authors have read and agreed to the published version of the manuscript. Funding: This research received no external funding. Acknowledgments: The authors would like to express their deepest gratitude to Alessandro Sabattini of Fincantieri S.p.A. and MAN Diesel and Turbo for supporting the present research. Particular thanks are due to Andrea Rossetto of Ranieri Tonissi S.p.A. Genova. Conflicts of Interest: The authors declare no conflict of interest.

Abbreviations List

AC Aftercooler B.m.e.p. Brake mean effective pressure C Turbocharger compressor DF Dual fuel Diff Difference E Economizer EG Electric generator EV Evaporator HFO Heavy fuel oil HRSG Heat recovery steam generator HWT Heat water tank ILV Isenthalpic lamination valve JW Jacket water MCR Maximum continuous rating MDO Marine diesel oil MFP Main feed pump NCR Normal continuous rating NG Natural gas Sat Saturated steam SC Scavenging system SCO steam condenser SCP Steam condenser pump SD Steam drum ST Steam turbine Superh Superheated steam T Turbocharger turbine, temperature WHR Waste heat recovery Energies 2020, 13, 985 20 of 21

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