Abstract

The considerable energy waste in maritime transport and the need to obtain alternatives to reduce emissions of polluting gases are factors that have motivated the study of waste heat recovery systems for marine engines. The system studied herein relies on a binary vapor cycle that uses water for the topping cycle while three organic fluids were investigated for the bottoming cycle: R601a, R134a, and R22. Each of these belongs to a different category of fluid, namely, dry fluid, isentropic fluid, and wet fluid, respectively. Two engines of different ratings and two different pressures of the heat recovery steam generator have been considered for each engine. Various outlet pressures for the topping turbine, which is the most liable to erosion and corrosion due to wet steam, have been investigated. The maximum efficiency achieved for the waste heat recovery system peaked at 21% while the maximum electric power accounted for 4.2% of engine brake power. Therefore, the employment of a waste heat recovery system based on a binary cycle seems a promising alternative to harnessing heat from the exhaust gases of marine engines.

Graphical Abstract Figure
Graphical Abstract Figure
Close modal

1 Introduction

To reduce the emission of greenhouse gas (GHG) produced by ships, several strategies have been investigated. Among the main approaches to reducing the ship fuel consumption are optimization of the hull and propeller shapes, hull coating, air injection under the hull, wind-assisted propulsion, and slow steaming [1]. Alternative low-carbon fuels have also been largely investigated, namely, liquefied petroleum gas and liquefied natural gas. For the low-speed two-stroke marine engines, which are the most efficient ones with about 50% efficiency, further in-cylinder improvements are not promising to reduce GHG emissions [2]. On the other hand, waste heat recovery systems (WHRS) could be considered an important alternative to produce more power with the same fuel consumption [3].

Heat dissipation occurs in different ways and at different temperatures in low-speed two-stroke marine engines, namely, by radiation, lubricating oil, jacket water, air cooler, and exhaust gases. The most significant heat waste occurs in the exhaust gases and at the highest temperature, being the most promising heat waste to be recovered. Generally, depending on the engine load and ambient conditions, the temperature of the gases leaving the turbocharger is from 250 to 300 C and this heat amount stands for about 50% of the total waste heat [4]. However, as the fuel burnt is usually heavy fuel oil (HFO), which contains sulfur, the risk of sulfuric acid formation and then corrosion in the exhaust stream limits the potential of heat recovery from this source [5].

The waste heat recovered can be used to heat the HFO and lubricating oil on board, and to produce fresh water, besides refrigeration and electric power to supply onboard consumers. The fuel-saving potential depends on the engine type, waste heat sources to be harnessed, system configuration, and the ship’s operational profiles, among others. With the purpose of power generation, a growing number of investigations have been performed on the conventional Rankine cycle (CRC), organic Rankine cycle (ORC), Kalina cycle (KC), and CO2-based power cycles on board ships. The application and feasibility of various WHRS for ships were discussed in Ref. [6], including thermoelectric generation and refrigeration systems. Multiple Rankine cycle types, KC, and combinations of technologies were addressed in Ref. [7]. The potential of using ORC as WHRS onboard ships was broadly addressed in Ref. [8], while a multi-objective optimization of a combination supercritical/transcritical CO2 WHRS was investigated in Ref. [9].

The CRC, which employs water as the working fluid, is a mature technology and has been largely applied to recover waste heat from the exhaust gases of engines. However, the high boiling temperature for low pressures of the water constrains the CRC efficiency. With the fluid flexibility of the ORC, working fluids with lower boiling points can be selected and this makes the ORC able to recover more energy from low-temperature waste heat sources [10]. The Kalina cycle takes advantage of the non-constant boiling temperature as it operates with ammonia–water mixtures while the cycles using pure substances as working fluids have constant boiling temperatures for fixed pressures [11]. Finally, CO2-based WHRS takes advantage of the moderate critical pressure and stability of this working fluid under high temperature and pressure, making it possible to employ transcritical and supercritical cycles [12,13].

Given such scenery, the present work employs two of the most mature technologies in a combined way and performs a thermodynamic study to find the best parameters combination. A binary vapor cycle composed of a CRC as the topping cycle and an ORC as the bottoming cycle is proposed as the WHRS for the exhaust gases from marine engines. Thus, the topping cycle where the source temperature is higher operates with water, which is a non-toxic, abundant, and low-cost working fluid, and three working fluids are explored for the bottoming cycle. Each of the fluids selected belongs to a different category, such as R601a (isopentane—C5H12) is a dry fluid, R134a (tetrafluoroethane—CH2FCF3) is an isentropic fluid, and R22 (chlorodifluoromethane—CHClF2) is a wet fluid [14]. Two engines of different ratings and two different pressures of the heat recovery steam generator (HSG) are also considered for each engine. Various outlet pressures for the topping turbine, which is the most liable to erosion and corrosion due to wet steam, are investigated and the combination of parameters providing the highest efficiency is found.

1.1 Background and the Proposed WHRS.

A binary vapor cycle combines two vapor cycles such that the condenser of the topping cycle (high-temperature cycle) serves as the boiler for the bottoming cycle (low-temperature cycle). A working fluid with advantageous high-temperature characteristics must be used in the topping cycle while another with advantageous low-temperature characteristics must be used in the bottoming cycle [15]. Thus, efficiency greater than each cycle has individually in its respective temperature range is expected for the binary cycle, according to Ref. [16].

Figure 1 shows the proposed binary vapor cycle arrangement as a WHRS for marine low-speed engines. Vapor is generated in the topping cycle by harnessing the waste heat from the engine exhaust gases in the HRSG. After that, vapor enters the turbine (state 1) and expands to the condenser pressure (state 2). The liquid that leaves the condenser (state 3) is pumped to the HRSG pressure (state 4). Meanwhile, the fluid in the bottoming cycle receives the heat rejected by the topping cycle when going through the boiler (condenser for the topping cycle). Therefore, the states a, b, c, and d for the bottoming cycle correspond to the states 1, 2, 3, and 4 in the topping cycle. However, the heat rejected by the bottoming cycle goes to the cooling water stream.

Fig. 1
Arrangement of a binary vapor cycle as WHRS
Fig. 1
Arrangement of a binary vapor cycle as WHRS
Close modal

Figure 2 illustrates the temperature–entropy diagram for a binary vapor cycle and what happens to the topping and bottoming cycles for different topping condenser pressures. The maximum and minimum temperatures (Tmax and Tmin) are dictated by the exhaust gas temperature and the environmental temperature, respectively. In both cases (Figs. 2(a) and 2(b)), those temperatures and the pinch-point temperature difference (ΔTpp) are kept the same. Therefore, as pressures for both topping and bottoming fluids are increased in the topping condenser (Fig. 2(b)), the pressure ratio decreases for the topping cycle and increases for the bottoming cycle. Since the efficiencies for both cycles are related to the temperature differential, which is in turn related to the pressure ratio, the topping cycle efficiency decreases as the bottoming cycle efficiency increases. It is also possible to see that the areas of the cycles have changed, that is, the net heat per unit mass of the cycles since the area equals net heat per unit mass in a temperature–entropy diagram. Hence, the net work per unit mass of each cycle has also changed because it equals the net heat per unit mass for reversible cycles. Additionally, the vapor quality at the turbine outlet in both cycles also varies for different topping condenser pressures. This is important because the liquid phase in the last stages of the turbine causes erosion, which increases maintenance costs and reduces the turbine’s life cycle.

Fig. 2
Temperature–entropy diagrams for different pressures in the topping cycle condenser: (a) lower pressures and (b) higher pressures
Fig. 2
Temperature–entropy diagrams for different pressures in the topping cycle condenser: (a) lower pressures and (b) higher pressures
Close modal

The effect of the topping condenser pressures over net work per unit mass and over vapor quality for each cycle depends on the saturation dome shape of the working fluids. In this context, there are three types of fluids: wet fluid, isentropic fluid, and dry fluid, as shown in Fig. 3. The main difference is the inclination of the saturation line on the vapor side. Wet fluids present a negative derivative of temperature to entropy (dT/ds) while isentropic fluids have an infinite derivative and dry fluids hold a positive derivative [17,18]. Thus, since a reversible process is considered, the fluid at the turbine outlet is a liquid–vapor mixture for wet fluids, saturated vapor for isentropic fluids, and superheated vapor for dry fluids.

Fig. 3
Saturation dome for different types of fluids
Fig. 3
Saturation dome for different types of fluids
Close modal

2 Methodology

All the computations were performed by a program developed by the authors in Octave [19] and using CoolProp [20] alongside, which is a library with pure and pseudo-pure fluid equations of state and transport properties for several substances. To ease the understanding of the methodology, three categories of variables were identified: parameters, decision variables, and dependent variables. The parameters are those remaining fixed throughout this study while the decision variables have different values explored, and dependent variables are those whose values are derived from other variables. Table 1 lists all the variables and their categories. To make the approach more realistic, apart from isentropic efficiencies, mechanical and electrical losses were also considered.

Table 1

Variables and their categories

Parameters
Efficiency of electric generators (ηEG=98%) and motors (ηEM=90%)
Efficiency of HRSG, bottoming boiler, and condenser (all taken as 100%)
Gearbox efficiency (ηGB=97%)
Isentropic efficiencies of turbines (ηisT=87%) and pumps (ηisP=83%)
Mechanical efficiencies of turbines (ηmT=99%) and pumps (ηmP=92%)
Minimum temperature (Tmin=Tc=35C)
Phase in the inlet of topping turbine (saturated steam in state 1)
Phase in the outlet of both condensers (saturated liquid in states 3 and c)
Pinch-point temperature difference (ΔTpp=15C)
Pressure loss in the pipes and heat exchangers (ΔpL = negligible)
Working fluid of the topping cycle (water)
Decision variables
Engine loads (25% EL 100%)
Engine models (8X52 and 12X92)
Pressure of steam leaving the HRSG (p1 = 7 bara and p1 = 9 bara, saturated vapor in both cases)
Steam quality at the outlet of the topping turbine (85% x 99% with step of 1%)
Working fluid for the bottoming cycle (R601a, R134a, and R22)
Dependent variables
Thermodynamic properties of interest in all states (1, 2, 3, 4, a, b, c, and d)
Mass flowrates for the topping and bottoming cycles (m˙T and m˙B)
Heat transfer rates in the condensers (Q˙Cd,T and Q˙Cd,B)
Electric power consumed by the motors (W˙EM,T and W˙EM,B)
Electric powers produced by the generators (W˙EG,T and W˙EG,B)
Efficiency for the topping and bottoming cycles (ηT and ηB)
Engine brake power (W˙B)
Parameters
Efficiency of electric generators (ηEG=98%) and motors (ηEM=90%)
Efficiency of HRSG, bottoming boiler, and condenser (all taken as 100%)
Gearbox efficiency (ηGB=97%)
Isentropic efficiencies of turbines (ηisT=87%) and pumps (ηisP=83%)
Mechanical efficiencies of turbines (ηmT=99%) and pumps (ηmP=92%)
Minimum temperature (Tmin=Tc=35C)
Phase in the inlet of topping turbine (saturated steam in state 1)
Phase in the outlet of both condensers (saturated liquid in states 3 and c)
Pinch-point temperature difference (ΔTpp=15C)
Pressure loss in the pipes and heat exchangers (ΔpL = negligible)
Working fluid of the topping cycle (water)
Decision variables
Engine loads (25% EL 100%)
Engine models (8X52 and 12X92)
Pressure of steam leaving the HRSG (p1 = 7 bara and p1 = 9 bara, saturated vapor in both cases)
Steam quality at the outlet of the topping turbine (85% x 99% with step of 1%)
Working fluid for the bottoming cycle (R601a, R134a, and R22)
Dependent variables
Thermodynamic properties of interest in all states (1, 2, 3, 4, a, b, c, and d)
Mass flowrates for the topping and bottoming cycles (m˙T and m˙B)
Heat transfer rates in the condensers (Q˙Cd,T and Q˙Cd,B)
Electric power consumed by the motors (W˙EM,T and W˙EM,B)
Electric powers produced by the generators (W˙EG,T and W˙EG,B)
Efficiency for the topping and bottoming cycles (ηT and ηB)
Engine brake power (W˙B)

Figure 4 illustrates the simulation tree followed to explore several system configurations. For the sake of clarity, not all topping turbine steam quality and engine load values are shown. Moreover, as the various values of engine load do not affect the efficiency and the net work per unit of mass, different engine load values were not simulated for different steam quality values. Therefore, the number of simulations totaled 324. It is worth noting that the engines investigated are of different numbers of cylinders (8 and 12) and different cylinder bores (52.0cm and 92.0cm), having one deemed small (8X52) and the other large (12X92) in the same category.

Fig. 4
Simulation tree to explore various system configurations
Fig. 4
Simulation tree to explore various system configurations
Close modal
Table 2 lists the decision variables related to the engine models chosen for the present study, as given by the GTD (General Technical Data) application.2 This application offers general technical data and guidelines for marine low-speed engines’ operation, installation, and arrangement. For both pressure values, the guiding steam production capacity considers saturated steam at those pressures, boiler efficiency of 100%, and feed water temperature of 80 C (state 4 in Fig. 2). However, as the steam quality at the outlet of the topping turbine assumes different values depending on the topping condenser pressure (states 2 and 3 in Fig. 2), the feed water also varies its temperature. Thus, the steam flowrate for the topping cycle (m˙T) is slightly different from the value given in Table 2 whenever T4 is different from 80 C. By considering that the same heat transfer rate occurs for various feed water temperatures, Eq. (1) was obtained to obtain the steam flowrate.
m˙T=m˙80C(h1h80Ch1h4)
(1)
Table 2

Decision variables related to the engine models

Engine load (%)Engine 8X52
Power (kW)Speed (rpm)Steam flowrate (kg/h)
7 bar9 bar
10014,48010528302280
9513,756103.222901770
9013,032101.417101210
8512,30899.51290810
8011,58497.51090630
7510,86095.4950500
7010,13693.2770340
60868888.6560180
50724083.3600270
40579277.4820550
30434470.3480270
25362066.1560370
Engine load (%)Engine 8X52
Power (kW)Speed (rpm)Steam flowrate (kg/h)
7 bar9 bar
10014,48010528302280
9513,756103.222901770
9013,032101.417101210
8512,30899.51290810
8011,58497.51090630
7510,86095.4950500
7010,13693.2770340
60868888.6560180
50724083.3600270
40579277.4820550
30434470.3480270
25362066.1560370
Engine load (%)Engine 12X92
Power (kW)Speed (rpm)Steam flowrate (kg/h)
7 bar9 bar
10077,4008013,93010,880
9573,53078.610,8907940
9069,66077.285005670
8565,79075.867003990
8061,92074.352702690
7558,05072.744702020
7054,1807140901770
6046,44067.542802260
5038,70063.554203720
4030,96058.973606020
3023,22053.664205340
2519,35050.460305100
Engine load (%)Engine 12X92
Power (kW)Speed (rpm)Steam flowrate (kg/h)
7 bar9 bar
10077,4008013,93010,880
9573,53078.610,8907940
9069,66077.285005670
8565,79075.867003990
8061,92074.352702690
7558,05072.744702020
7054,1807140901770
6046,44067.542802260
5038,70063.554203720
4030,96058.973606020
3023,22053.664205340
2519,35050.460305100

Table 3 lists the main features of the fluids investigated in the bottoming cycle whereas only water was used in the topping cycle. Reasonable differences can be seen in the features of the fluids except regarding the safety group, given by the ANSI/ASHRAE (American National Standards Institute/American Society of Heating, Refrigerating and Air-Conditioning Engineers) Standard [21], and the ozone depletion potential (ODP). Both fluids R134a and R22 belong to the no-flame-propagation category (assigned by the number 1), and R601a is categorized as a higher flammability fluid (assigned by the number 3) whereas all of them are categorized as lower toxicity fluids (assigned by the letter A). Regarding ODP, both fluids, R601a and R134a, hold no potential to deplete the ozone layer while the release of 1.0 kg of R22 is equivalent to 0.05 kg of R11 (trichlorofluoromethane—CCl3F), which is used as the reference [22]. The global warming potential for a 100-years-timeline (GWP100) is given as a measure of equivalent kilograms of CO2, this means, for instance, that the impact caused by 1.0 kg of R601a is equivalent to 3.0 kg of CO2 [23].

Table 3

Main features of the fluids for the bottoming and topping cycles

FeatureBottoming cycleTopping cycle
R601aR134aR22Water
Critical pressure (bara)33.7840.5949.9220.6
Critical temperature (C)187.2101.196.15374
Saturation dome typeDryIsentropicWetWet
Safety groupA3A1A1
Ozone depletion potential (ODP)000.05
Global warming potential (GWP100)314301810
FeatureBottoming cycleTopping cycle
R601aR134aR22Water
Critical pressure (bara)33.7840.5949.9220.6
Critical temperature (C)187.2101.196.15374
Saturation dome typeDryIsentropicWetWet
Safety groupA3A1A1
Ozone depletion potential (ODP)000.05
Global warming potential (GWP100)314301810
The dependent variables were obtained based on the parameters, decision variables, and the equations derived from energy balances in every system component. For such balances, the effects of kinetic and potential energies were neglected as they are small relative to other energy transfers. Therefore, the mass flowrate for the bottoming cycle (m˙B) is given by Eq. (2), while the mass flowrate for the topping cycle (m˙T) was already given by Eq. (1). The heat transfer rates in the HRSG (Q˙T) and the topping condenser/bottoming boiler (Q˙B) are, respectively, given by Eqs. (3) and (4). Notice that h (subscripts 1–4 and a–d) denotes specific enthalpy for the various thermodynamic states in Fig. 2.
m˙B=m˙T(h2h3hahd)
(2)
Q˙T=m˙T(h1h4)
(3)
Q˙B=m˙B(hahd)
(4)
The electric power consumed by the topping and bottoming motor pumps (W˙EM,T and W˙EM,B) are, respectively, given by Eqs. (5) and (6).
W˙EM,T=m˙T(h4h3)ηmPηEM
(5)
W˙EM,B=m˙B(hdhc)ηmPηEM
(6)
The electric powers produced by the topping and bottoming turbo-generators (W˙EG,T and W˙EG,B) are, respectively, given by Eqs. (7) and (8).
W˙EG,T=m˙T(h1h2)ηmTηGBηEG
(7)
W˙EG,B=m˙B(hahb)ηmTηGBηEG
(8)
Therefore, the net electric power for the topping and bottoming cycles (W˙net,T and W˙net,B) are calculated by Eqs. (9) and (10), respectively, while the net electric power for the binary vapor cycle (W˙net) is calculated by Eq. (11).
W˙net,T=W˙EG,TW˙EM,T
(9)
W˙net,B=W˙EG,BW˙EM,B
(10)
W˙net=W˙net,T+W˙net,B
(11)
The efficiency for the topping and bottoming cycles (ηT and ηB) are calculated individually by Eqs. (12) and (13), respectively, while the efficiency of the binary vapor cycle (η) is calculated by Eq. (14).
ηT=W˙net,TQ˙T
(12)
ηB=W˙net,BQ˙B
(13)
η=W˙net,T+W˙net,BQ˙T
(14)

2.1 Verification of the Methodology.

To verify the accuracy of the methodology, a binary vapor power cycle, consisting of steam and R134a as the working fluids, was simulated and compared with an analytical sample from the literature [16]. In such a sample, a 2.0 kg/s mass flowrate of steam enters the topping turbine at 8 MPa and 600 C, and saturated liquid exits the topping condenser at 250 kPa. Superheated R134a leaves the bottoming evaporator at 600 kPa and 30 C, and saturated liquid leaves the bottoming condenser at 100 kPa. Table 4 lists the values of some features computed with the present approach and their values from the literature as well as the relative deviations. One can see that the highest deviation peaked at under 1%, thus the approach is considered accurate enough.

Table 4

Verification of the developed approach accuracy

FeatureComputedLiteratureRelative deviation (%)
Power of the topping turbine (kW)1876.0821876.00.004
Power of the topping pump (kW)16.54216.50.255
Heat transfer rate of the topping steam generator (kW)6197.5246196.70.013
Bottoming mass flowrate (K/s)17.13617.30.948
Power of the bottoming turbine (kW)631.393635.10.584
Power of the bottoming pump (kW)6.2206.20.323
Efficiency of the binary vapor power plant (%)40.09240.20.269
FeatureComputedLiteratureRelative deviation (%)
Power of the topping turbine (kW)1876.0821876.00.004
Power of the topping pump (kW)16.54216.50.255
Heat transfer rate of the topping steam generator (kW)6197.5246196.70.013
Bottoming mass flowrate (K/s)17.13617.30.948
Power of the bottoming turbine (kW)631.393635.10.584
Power of the bottoming pump (kW)6.2206.20.323
Efficiency of the binary vapor power plant (%)40.09240.20.269

3 Results and Discussion

The quality of the steam leaving the topping turbine depends on the topping condenser pressure. Since a constant pinch-point temperature difference was assumed, the pressure of the bottoming fluid leaving the topping condenser (bottoming evaporator) is also related to the topping turbine steam quality. Therefore, Fig. 5 illustrates such a relationship for both values of steam pressure in the HRSG and the three bottoming fluid alternatives. The different curve shapes are due to maximum pressure values being limited to the critical pressure of every refrigerant. One can see that the refrigerant with the lowest pressure is R601a and it is followed by far for R134a and R22. For instance, given the steam quality of 90% for 9 bar HRSG pressure, R601a would be at 5.1 bar while R134a and R22 would be, respectively, at 29.2 bar and 40.3 bar. Since it is riskier and usually more expensive to have higher pressures in the system, selecting R601a could be advantageous.

Fig. 5
Pressures in the topping condenser for different steam pressures and steam quality values at the topping turbine outlet
Fig. 5
Pressures in the topping condenser for different steam pressures and steam quality values at the topping turbine outlet
Close modal

Figure 6 shows the behavior of the efficiency for the bottoming, topping, and binary cycles with various values of steam quality. Since the HRSG pressure is given, the difference of the WHRS installed in different engines is only the steam flowrate and this does not affect the efficiency, thus the curves are the same regardless of the engine model. The general behavior of curves is observed regardless of the steam pressure and the bottoming fluid: as the steam quality increases, the efficiency of the topping cycle decreases linearly while the efficiency of the bottoming cycle increases. This renders an optimum value of quality that maximizes the efficiency of the binary vapor cycle. The refrigerant with the best performance was R601a, with maximum efficiency of 20.3% for 7 bar (88% quality) and 21.5% for 9 bar (87% quality), while the others are equivalent to each other from an efficiency point of view. Those efficiencies may be thought low but considering that the maximum temperature for the 7-bar and 9-bar cycles are, respectively, 165 C and 175 C (water saturation temperatures), their Carnot efficiencies are 29.7% and 31.3%. Notice that the disturbances occurring around 92% quality for R22 and more intensely for R134a are due to the previously mentioned pressure limit. The curvatures of the saturation domes of R22 and R134a around their critical points cause those slight drops.

Fig. 6
Efficiency variation for different steam pressures with various steam quality values at the topping turbine outlet: (a) both engines with 7 bar steam pressure and (b) both engines with 9 bar steam pressure
Fig. 6
Efficiency variation for different steam pressures with various steam quality values at the topping turbine outlet: (a) both engines with 7 bar steam pressure and (b) both engines with 9 bar steam pressure
Close modal

The electric power that could be harvested from the engines’ exhaust gases by the proposed WHRS follows the same trends seen for efficiency, as shown in Fig. 7. However, in this case, the HRSG pressure of 7 bar leads to a higher power while leading to a slightly lower efficiency (Fig. 6). This is explained by the considerably higher mass flowrates of steam produced by the HRSG when it works at 7 bar (Table 2). For the lower values of steam quality, all refrigerants present quite similar performance while R22 and R134a are approximately equivalent to each other all over the range of study. The refrigerant with the best performance was R601a, leading to a maximum power of 386.2 kW for 7 bar (89% quality) and 331.3 kW for 9 bar (88% quality), considering engine 8X52, and leading to a maximum power of 1903 kW for 7 bar (89% quality) and 1582 kW for 9 bar (88% quality), considering engine 12X92. This means that there is a gain related to the use of 7 bar accounting for 16.6% and 20.3% for the small engine and the large engine, respectively. It is worth saying that Fig. 7 considers engine loads as 100%.

Fig. 7
Variation of generated power for different engines and steam pressures with various steam quality values at the topping turbine outlet: (a) engine 8X52 with 7 bar steam pressure, (b) engine 8X52 with 9 bar steam pressure, (c) engine 12X92 with 7 bar steam pressure, and (d) engine 12X92 with 9 bar steam pressure
Fig. 7
Variation of generated power for different engines and steam pressures with various steam quality values at the topping turbine outlet: (a) engine 8X52 with 7 bar steam pressure, (b) engine 8X52 with 9 bar steam pressure, (c) engine 12X92 with 7 bar steam pressure, and (d) engine 12X92 with 9 bar steam pressure
Close modal

Figure 8 highlights the benefit of employing a binary vapor WHRS instead of using the conventional Rankine cycle alone, that is, using only the topping cycle. The surplus electric power, given as a percentage, is here used to compare the net electric power produced by the binary vapor cycle with that produced by the topping cycle alone. One can see that the benefit of adopting a binary cycle increases as the steam quality increases (higher topping condenser pressures), regardless of the pressure. Notice that the optimum quality (ranging from 87% to 89%, depending on the refrigerant) is the one to maximize the net electric power of the binary cycle but does not represent the point with the largest contribution. Using a binary vapor cycle water-R601a renders a surplus power of 54.8% and 64.4% when compared to the topping cycle alone respectively for 7 bar and 9 bar and a steam quality of 90%, which is generally the minimum quality adopted for the sake of maintenance. This could be taken as encouraging regarding the use of binary vapor cycles, but economic aspects must be considered to analyze the actual viability of the system.

Fig. 8
Surplus power due to the bottoming cycle for different steam pressures with some steam quality values at the topping turbine outlet: (a) both engines with 7 bar steam pressure and (b) both engines with 9 bar steam pressure
Fig. 8
Surplus power due to the bottoming cycle for different steam pressures with some steam quality values at the topping turbine outlet: (a) both engines with 7 bar steam pressure and (b) both engines with 9 bar steam pressure
Close modal

To verify how representative it is the net electric power that can be harnessed from the exhaust gases by a binary vapor cycle, the power ratio (net electric power over engine brake power) is shown in Fig. 9. One can see quite different trends regarding the two engine models; the highest power ratios occur in the highest loads for engine 8X52 and the lowest loads for engine 12X92, peaking, respectively, at 2.7% (100% load, 7 bar, and R601a) and 4.2% (25% load, 7 bar, and R601a). In both cases, the lowest power ratios occur in intermediate loads, having 60% and 70%, respectively, for engines 8X52 and 12X92. This is explainable by the fact that such engines are the most efficient around those values of load, so there is less energy to be harnessed in the exhaust gases. The different refrigerants do not prove to be very influential, with R22 and R134a having overlapping curves and R601a being slightly better. On the other hand, the steam pressure of 7 bar renders power ratios considerably larger for both engines. It is worth saying that Fig. 9 considers 90% quality for the steam leaving the topping turbine. Moreover, the change in the curves for loads lower than 40% may be attributed to the necessity for the blower operation, which affects the exhaust temperature and mass flowrate, changing the steam production.

Fig. 9
Variation of power ratio for different engines and steam pressures with various engine loads: (a) engine 8X52 and (b) engine 12X92
Fig. 9
Variation of power ratio for different engines and steam pressures with various engine loads: (a) engine 8X52 and (b) engine 12X92
Close modal

4 Conclusion

The present work proposed a WHRS for marine engines’ exhaust gases based on a binary vapor cycle composed of a CRC as the topping cycle and an ORC as the bottoming cycle. Three refrigerants with different saturation domes and thermophysical properties were investigated. Two engines and two pressure values for the steam produced by the HRSG were also investigated. Moreover, various outlet pressures for the topping turbine, which lead to different steam qualities, were investigated.

Although each of the refrigerants belongs to a different category and holds distinct properties, R601a performed only slightly better than the others. Fluids R22 and R134a presented coinciding performances. Since R22 is already phased out because of its ODP, R134a has proven to be a substitute without such an issue. However, R134a has a high GWP and R601a has high flammability, such that other refrigerants are advisable to be addressed in future works.

The use of a bottoming cycle increased the electric net power by up to 64% and the efficiency of the binary vapor cycle reached over 20%. The optimum steam quality has laid slightly under 90% for all refrigerants, such that there is no relevant loss in adopting 90%, and over 4% of the engine brake power could be produced as net electric power by the binary cycle. The pressure of 7 bar for the HRSG has proven to render substantial gains. In this sense, pressures even lower are advisable to be addressed in future works. Furthermore, although the proposed WHRS has shown to be promising from an energetic perspective, an exergoeconomic study should be addressed in future work to prove the viability of such a system.

Footnote

Acknowledgment

J.A.S. thanks CNPq for financial support (process: 306024/2017-9) and C.H.M. thanks FAPERGS for financial support to the ARD research project (19/2551-0001250-0).

Conflict of Interest

There are no conflicts of interest.

Data Availability Statement

The authors attest that all data for this study are included in the paper.

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