1 Introduction

The search for renewable energy sources and increasing energy efficiency in power generation has been a common concern in recent years. In addition, there is a concern over emissions of greenhouse gases such as CO2 from fossil fuel combustion. Some studies aim at increasing efficiency and consequently reducing the usage of fossil fuel to generate electricity. Most thermal power plants are based on Rankine cycles, gas turbines (Brayton cycle), alternative engines (otto/diesel) or a combination of these technologies. The efficiency of these systems may be increased using absorption chillers or organic Rankine cycles (ORC) powered by the exergy remaining in the heat wasted by these power plants.

The use of cooling equipment to enhance the power generation is extensively present in the literature. The performance of gas turbines for example is dependent on ambient temperature [15]. The most common application for cooling equipment to enhance power production is reducing the temperature of gas turbine inlet air to increase the air mass flow rate through the equipment and also to reduce compressor power consumption [6]. Shirazi et al. [22] presented a mathematical model for an ice thermal energy storage (ITES) system applied to cool the gas turbine inlet air. It was concluded that using ITES, an improvement of 11.63% in power output and 3.59% in the exergy efficiency can be achieved. The efficiency of power plants based on Rankine cycle depends on the condensation temperature. A reduction in this temperature results in reduction of condensation pressure, which means more power output during the expansion in the steam turbine and consequently greater plant efficiency. A particular plant may have an increase of around 10% in power generation when the condenser pressure is changed from 0.087 to 0.033 bar [2]. Munõz et al. [18] analyzed the performance of a Rankine cycle in which the energy from boiler exhausting gases was used to feed an absorption chiller to chill the condenser cooling water. The conclusion was that the energy of exhausting gases was not enough to significantly reduce the temperature of cooling water due to the high mass flow rate of this stream. In the present work, the use of a cold water tank to store the water cooled by the absorption chiller is proposed. This water will be used during hours of peak consumption to increase the power produced by the power plant.

Recently, ORCs have been studied due to their ability to use low-grade heat sources such as waste heat from industry, geothermal energy and solar energy to produce electricity [4]. ORCs are more interesting than conventional Rankine cycles, which use water/steam as a working fluid, due to their physical properties more adequate to deal with low temperature heat sources. Some of the important properties are the positive inclination of vapor saturation line and proper condensation and evaporation pressure under the temperatures of the environment and the heat sources, respectively. The use of organic fluids also results in more compact system and reduces the costs of implementation and maintenance of the system [20]. Song et al. [26] analyzed an ORC to recover waste heat from marine diesel engines using jacket cooling water and the exhaust gases to obtain maximum power output. The cycle that used the jacket cooling water to pre-heat the working fluid and the exhaust gases to evaporate it resulted in a gain of 99.7 kW in power generation with low implementation costs. Galindo et al. [11] analyzed thermodynamically and experimentally an ORC using waste heat from a gasoline engine. Other authors such as Shu et al. [23] and Battista et al. [3] studied the effects of the operational characteristics on the performance of ORCs attached to combustion engines. The first concluded that azeotropic mixtures of organic fluids result in lower efficiencies than most pure organic fluids, while zeotropic mixtures, in certain proportions, result in higher efficiency when compared to the same pure organic fluids. The second concluded that the installation of an ORC coupled to a conventional gasoline engine with turbo-compressor can increase the mechanical efficiency of the system in up to 3.7%.

Chacartegui et al. [7] analyzed the applicability of ORC in using the exhaust gases from regenerative gas turbines. This configuration compared to conventional gas turbines results in exhaust gases at lower temperatures which can be used as heat source for the ORC. The system optimization was conducted based on gas turbine inlet temperature and resulted in a gain of efficiency superior to those from the combined cycles presented in the literature. Clemente et al’s. [9] study was focused on the application of an ORC to use exhaust gases from small scale gas turbines (100 kWe). Comparing the options for the ORC expanders, the author concluded that single stage turbines are more advantageous than positive displacement expanders because they can be connected to the gas turbine shaft and produce higher gains in efficiency to the whole system. For this application, micro-turbines, refrigerant fluids are more suitable because they allow the use of compact equipments. Camporeale et al. [5] analyzed thermodynamically and technical economically an ORC using waste heat from a non-regenerative gas turbine powered by biomass. The absence of a regenerator results in higher temperature of the exhaust gases. It was concluded that superheating on subcritical ORCs causes efficiency losses. Therefore, it is preferable to use cycles with evaporation close to the saturation line. The same issue was observed for supercritical cycles. In this case, the parameters should be kept next to the critical point. For this specific application, the most suitable fluid is the toluene due to its high heat recovery rate. Chacartegui et al. [8] studied the application of an ORC to use exhaust gases from a humid air turbine for distributed micro-generation. Results show a gain of 25% in power generated which corresponds to an increase of 11% in the efficiency when compared to an isolated gas turbine. The economical analysis showed a reduction of 15% on the cost of electricity which indicates that this technology is highly competitive for the proposed application. Demierre et al. [10] designed a prototype of a thermally driven heat pump, which consisted in two ORC coupled, one generating shaft power to rotate a compressor of a reversed ORC which worked as the heat pump. The upper ORC consisted in a supercritical ORC which used R134a as the working fluid. The hot heat source used is the combustion gases from stoichiometric combustion of methane. The cold heat source is the brine from a geothermal probe. The results of the experiment are compared to the simulations and resulted on a difference of less than 3% between the predicted overall heat transfer and the measurements.

In this work, the ORC makes use of the opportunity created when the fuel used in a boiler is replaced by a fuel with lower sulfur concentration, therefore, the energy remaining in exhausting gases can be further used without reaching the acid condensation point. An optimization procedure is performed to find the configuration which provides maximum power output for each tested fluid. The result of the best ORC scheme is compared to the result from using energy storage and an absorption chiller to chill condenser cooling water.

2 System modeling

For this work, a typical power plant based on Rankine cycle is used. The following sub topics present some simplifying assumptions, the characteristics of the power plant and the proposed changes to increase both power generation and plant efficiency.

2.1 Simplifying assumptions

  1. 1.

    Changes in potential and kinetic energies are considered negligible.

  2. 2.

    All components operate under steady state condition except for the water tanks.

  3. 3.

    The pressure drop in pipelines is disregarded.

  4. 4.

    Environment temperature and humidity remains constant along the analysis.

2.2 Studied Rankine cycle-based power plant

This analysis is carried out considering the power plant shown in Fig. 1.

Fig. 1
figure 1

Studied power plant

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The boiler produces 27.78 kg/s (100 ton/h) of steam at 485 °C and 9100 kPa. These data follow those from the conventional boilers installed in the biggest Brazilian refinery [24, 25]. Exhaust gases leave at 175.6 °C. Saturated steam enters into the condenser at 45 °C. The water exits the cooling tower at 30 °C and returns from condenser at 40 °C. The temperature and pressure of the subcooled liquid leaving the deaerator are, respectively, 133 °C and 300 kPa. All pumps have isentropic efficiency of 80%. The steam turbine has isentropic efficiency of 85% and the boiler has thermal efficiency of 92%. Tables 1 and 2 show, respectively, the properties of each stream of the plant and the global information.

Table 1 Studied power plant thermodynamic properties
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Table 2 Global information of the studied power plant
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2.3 Exhausting gas acid dew point

According to Ganapathy [12], if the temperature of the exhausted gases reaches the acid dew point temperature, corrosion problems may occur. It is a general trend, nowadays, the replacement of sulfur rich fuels by fuels with lower concentration of this element to avoid excessive SO x emissions. The most well-known example in Brazil is the gradual replacement of S1800 diesel by S10 diesel during the last decades [1, 19]. This kind replacement also allows further use of exhausting gas energy because the dew point temperature will be lowered as the sulfur concentration decreases. Table 3 shows some commonly used fuel, their composition and dew point. Fuel oil 1A was selected due to its lower amount of sulfur as the substitute to the originally used fuel, a refinery fuel oil (not commercially available), which allows 175.6 °C as exhausting temperature.

Table 3 Table 3 Fuels studied in this work. Adapted from [21]
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2.4 First proposal: absorption chiller

In this proposal an absorption chiller is coupled to the boiler of the power plant to make use of boiler exhaust gases as heat source. During the operation of the plant, the chiller cools water and directs it into the cold water storage tank. At peak electricity consumption hours, the stored cold water is drained into the condenser, replacing the cooling water from cooling tower. The cold water stored reduces the condensation pressure increasing the enthalpy drop through the steam turbine and resulting in higher power output and efficiency. The selected absorption chiller has COP of 0.72, a nominal cooling capacity 400 ton with 60.56 kg/s of maximum mass flow rate. The chilled water temperature is 6.67 °C and the chiller electrical consumption is 7.2 kW. Figure 2 shows the proposed system.

Fig. 2
figure 2

Proposed power plant with absorption chiller

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2.4.1 Standard operation mode

Under standard mode, the power plant operates according to Fig. 1. The condenser is cooled by the cooling tower water and the absorption chiller cools the water stored in the hot water storage tank. This cold water, after being cooled by the chiller, is stored into the cold water storage tank. When the cold water tank is full, the power plant is ready to operate under repowering mode.

2.4.2 Repowering operation mode

Under repowering mode, the water flow from the cooling tower is bypassed and the cold water from the tank is drained into the condenser and returns to the hot water storage tank. The mass flow rate needed to cool the condenser is considerably greater than the mass flow rate of cold water that the absorption chiller can provide. Therefore, this repowering operation mode only occurs for a short period of time. Note, that a single stratified tank (with a thermocline) could also be used. Since the absorption chiller is just a component of the overall system studied, a black box approach has been used to model this component.

2.4.3 Determination of the repowering operation time

During repowering mode, the cold water is drained into the condenser with mass flow rate \(\dot{m}_{\text{resf}}\), while the cold water storage tank receives water cooled by the chiller with mass flow rate \(\dot{m}_{\text{chiller}}\). The length of time in which the repowering mode t R occurs is determined by:

$$M_{\text{water}} = (\dot{m}_{\text{resf}} - \dot{m}_{\text{chiller}} )t_{\text{R}} ,$$
(1)

M water is the tank capacity in mass (kg). After the full discharge of cold water, the cycle returns to operation in standard mode during a period of time t S.

$$M_{\text{water}} = \dot{m}_{\text{chiller}} t_{\text{S}} .$$
(2)

For the proposed system, a tank with a capacity of 2915.29 m3 is required. Considering that the tank diameter is equal to its height both values are 15.48 m. The total power generated during the repowering operation mode is given by:

$$\dot{W}_{\text{repowered}} = \dot{W}_{\text{plant}} - \dot{W}_{\text{chiller}} - \dot{W}_{\text{pump}} ,$$
(3)

\(\dot{W}_{\text{plant}}\) is the power produced by power plant, \(\dot{W}_{\text{chiller}}\) is the electrical power consumed by the chiller and \(\dot{W}_{\text{pump}}\) is the power consumed by cold water pump.

2.4.4 Condenser analysis

During standard operation, the heat rejected by the condenser \(\dot{Q}_{\text{rej}}\) is determined by Eqs. (4) and (5):

$$\dot{Q}_{\text{rej}} = U_{\text{C}} A_{\text{C}} \Delta T_{\text{ml}} ,$$
(4)
$$\dot{Q}_{\text{rej}} = \dot{m}_{\text{w}} \Delta h_{\text{Cond}} ,$$
(5)

U C is the overall heat transfer coefficient considered as 2.5 kW/m2 K [13]. A C is the total area of heat exchange. \(\dot{m}_{\text{w}}\) is the mass flow rate of working fluid condensing. \(\Delta h_{\text{Cond}}\) is the working fluid enthalpy variation in the condenser. \(\Delta T_{\text{ml}}\) is the logarithmic mean temperature difference of the streams in the condenser. The calculated area of heat exchange for the condenser is 2068 m2. Figure 3a shows the arrangement for the condenser during standard operation mode.

Fig. 3
figure 3

Arrangement for the condenser: a standard operation mode and b repowering operation mode

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During repowering operation, the heat rejected by the condenser is calculated iteratively, as indicated at Table 4, by Eqs. (6)–(8). Equation (9) is the objective function that must be zero. \(\dot{m}_{{{\text{w}}\_{\text{rep}}}}\) is the working fluid mass flow rate for the repowering operation mode, C p is the specific heat of the water and \(\Delta T_{23 - 20}\) is the temperature difference between streams 20 and 23 shown in Fig. 3b.

$$\dot{Q}_{{{\text{rej}}1}} = U_{\text{C}} A_{\text{C}} \Delta T_{\text{ml}} ,$$
(6)
$$\dot{Q}_{{{\text{rej}}2}} = \dot{m}_{{{\text{w\_rep}}}} \Delta h_{\text{Cond}} ,$$
(7)
$$\dot{Q}_{{{\text{rej}}3}} = \dot{m}_{\text{resf}} C_{\text{p}} \Delta T_{23 - 20} ,$$
(8)
$$D_{\text{iff}} = \sqrt[2]{{(\dot{Q}_{{{\text{rej}}1}} - \dot{Q}_{{{\text{rej}}2}} )^{2} + (\dot{Q}_{{{\text{rej}}1}} - \dot{Q}_{{{\text{rej}}3}} )^{2} }}.$$
(9)
Table 4 Parameters used for iterative calculation for the heat rejected by the condenser
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2.4.5 Results

During repowering operation, the condensation temperature decreases resulting in less heat rejected to the ambient. Therefore, the plant efficiency and power generation are increased. A reduction of 2.24% in the steam quality, x steam, at steam turbine exit is noticed. The lower value of steam quality may increase the maintenance requirements due to the erosion caused by the impact of water droplets on the blades of the turbine low pressure stages. Table 5 shows a comparison between the condenser parameters during standard and repowering operation. Table 6 shows the calculated global data for the power plant in repowering operation mode. Table 7 shows the benefits of the use of the absorption chiller to the studied plant.

Table 5 Condenser parameter comparison
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Table 6 Calculated global data during repowering operation
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Table 7 Main results
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2.4.6 System optimization

Different ratios of cooling water from cooling tower and the cold water, cooled by the absorption chiller, were mixed and tested to find the best ratio to increase power generation. A higher temperature mixture enables a longer period of repowering. On the other hand, lower temperatures last for short periods but the increase in power production is more accentuated. Figure 4 shows the daily energy gain according to the percentage of chilled water mixed with cooling tower water. The maximum increment in energy is obtained using 100% of chilled water for a shorter period of time instead of mixed water for a longer period of time. Therefore, the results previously presented correspond to the optimized ratio.

Fig. 4
figure 4

Energy gain according to the percentage of chilled water in the mixture

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2.5 Second proposal: organic Rankine cycle

In this section, the ORC configurations, thermodynamic model and methodology are described. Both recuperative ORC (RORC) and non-recuperative ORC (NORC) were tested. Supercritical conditions were used for fluids which perform better in these conditions.

2.5.1 Non-recuperative ORC

Figure 5 shows the possible configurations for the NORC in the T–s diagrams. Subcritical conditions are shown in Fig. 5a. The fluid enters the turbine with temperature and pressure defined in state 14. Then it expands to the condensation pressure (state 15) and enters the condenser, leaving at the same pressure but as saturated liquid. This liquid is pumped to the evaporation pressure (state 17), which is lower than the critical pressure, and enters the economizer, where it is heated at constant pressure to saturated liquid state (state 24). Next, through the evaporator, the fluid evaporates and leaves as saturated vapor (state 25). If it is advantageous for the cycle, the fluid is superheated next, reaching state 14. Otherwise, state 14 is the same as state 25. The heating process from state 17 to state 14 is obtained using the waste heat source (boiler exhausting gases) that goes from state 12 to state 13 as shown in Fig. 5a.

Fig. 5
figure 5

Configuration and T–s diagram of NORC: a subcritical and b supercritical

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Figure 5b shows the supercritical configuration for the NORC. The difference between both configurations is that the fluid enters the turbine in a supercritical condition. That means that the temperature and pressure of the fluid are above critical values. Therefore, there is no evaporation process; the supercritical fluid goes directly from state 17 to state 14, above saturation line. Fluids that can operate in supercritical condition with the given heat source temperature were tested to confirm if they perform better that way.

2.5.2 Recuperative ORC

This configuration is represented in Fig. 6. The reason to use this layout is because some fluids leave the turbine after expansion as superheated vapor. The use of a recuperator allows transferring heat from the superheated vapor to the fluid leaving the pump. The condition for applying this layout is that T 15 > T 17. The fraction of the heat recovered is shown in Fig. 6 as area 1. This heat would originally be lost in condenser to the ambient. Similar to the NORC, the RORC can also be either subcritical (Fig. 6a) or supercritical (Fig. 6b) depending on same factors exposed in Sect. 2.5.1.

Fig. 6
figure 6

Configuration and T–s diagram of RORC: a subcritical and b supercritical

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2.5.3 Thermodynamic model

To compare the results of the proposals, the heat source characteristics for the ORC are the same used for the absorption chiller. The exhaust gas mass flow rate is 35.73 kg/s at 175.61 °C. The specific heat used was the same used for the absorption chiller (1.087 kJ kg−1 K−1). The minimum temperature at state 13 was limited to 5 °C above the acid dew point at 106.5 °C. All components were modeled at steady state and the same assumptions from Sect. 2.1 were applied. Table 8 gives the parameters used for the simulations. Some of these parameters were set according to Mazetto et al. [17].

Table 8 ORC specifications for simulation
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The isentropic efficiencies of the expander and pump are given, respectively, in Eqs. (10) and (11).

$$\eta_{{t,{\text{ise}}}} = \frac{{h_{14} - h_{15} }}{{h_{14} - h_{{15,{\text{ise}}}} }},$$
(10)
$$\eta_{\text{p,ise}} = \frac{{h_{{17,{\text{ise}}}} - h_{16} }}{{h_{17} - h_{16} }}.$$
(11)

In these equations h is the specific enthalpy at the subscripted states. The energy balances for the expander and pump are given, respectively, in Eqs. (12) and (13).

$$\dot{W}_{\text{t}} = \dot{m}_{\text{wf}} ({\text{h}}_{14} - h_{15} ),$$
(12)
$$\dot{W}_{\text{p}} = \dot{m}_{\text{wf}} ({\text{h}}_{17} - h_{16} ),$$
(13)

\(\dot{m}_{\text{wf}}\) is the mass flow rate of working fluid. Equations (14)–(17) give the energy balances of the economizer, evaporator, super-heater and supercritical evaporator.

$$\dot{m}_{\text{g}} C_{{{\text{p}}_{\text{g}} }} (T_{22} - T_{13} )= \dot{m}_{\text{wf}} ({\text{h}}_{24} - h_{17} ),$$
(14)
$$\dot{m}_{\text{g}} C_{{{\text{p}}_{\text{g}} }} (T_{21} - T_{22} ) = \dot{m}_{\text{wf}} ({\text{h}}_{25} - h_{24} ),$$
(15)
$$\dot{m}_{\text{g}} C_{{{\text{p}}_{\text{g}} }} (T_{12} - T_{21} ) = \dot{m}_{\text{wf}} ({\text{h}}_{14} - h_{25} ),$$
(16)
$$\dot{m}_{\text{g}} C_{{{\text{p}}_{\text{g}} }} (T_{12} - T_{13} ) = \dot{m}_{\text{wf}} ( {\text{h}}_{14} - h_{17} ).$$
(17)

The recuperator effectiveness is defined in Eq. (18). Energy balance for this equipment is given by Eq. (19)

$$\varepsilon_{\text{r}} = \frac{{T_{15} - T_{x} }}{{T_{15} - T_{17} }},$$
(18)
$$h_{15} - h_{x} = h_{y} - h_{17} .$$
(19)

The heat transferred from the exhaust gases to the working fluid is calculated by Eq. (20). Equations (21) and (22) give the net power generated and cycle efficiency.

$$\dot{Q}_{\text{H}} = \dot{m}_{\text{wf}} ({\text{h}}_{14} - h_{17} ),$$
(20)
$$\dot{W}_{\text{net}} = \dot{W}_{\text{t}} - \dot{W}_{\text{p}} ,$$
(21)
$$\eta_{\text{ORC}} = \frac{{\dot{W}_{\text{net}} }}{{\dot{Q}_{\text{H}} }}.$$
(22)

2.5.4 Methodology for optimization of the ORCs

To find the configuration for the highest net power output, 23 fluids were tested. Table 9 shows the fluid candidates and their respective properties. The selected fluids comply with Kyoto and Montreal protocols and are the same used by Mazetto et al. [17]. The software EES® [16] was used for modeling the cycle. This software has a large database of fluid properties and is well known in scientific media. The optimization method employed was the Genetic Algorithm which consists on a computational model inspired by Darwin’s theory of evolution in search for the global optima. At the first moment, the software calculates the objective function for a population of random values for the variables. After that, a new generation is chosen taking into account the best previous results and new results are calculated. After some generations, it is possible to find the optimal values for these variables. To avoid local maxima, the method also applies mutations in each generation, which are values far from the population tested, to find the global maximum.

Table 9 Working fluid candidate properties
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To maximize the objective function (net power output, \(\dot{W}_{\text{net}}\)) given by Eq. (21), the optimized variables were T 14 and P 14, respecting the constraint parameters from Table 8: pinch temperature difference at evaporator and minimum temperature at state 13 (T 13). For the subcritical ORCs, the maximum turbine inlet temperatures and pressures allowed were set to be the critical temperatures and pressures of each fluid tested. For the supercritical states, the critical temperatures and pressures were set as the minimum turbine inlet temperatures and pressures, respectively.

2.5.5 Results

The best performance fluids for non-recuperative ORC are presented in Table 10. The best performance fluids for recuperative ORC are presented in Table 11. Both tables present, for each substance, the optimized values for the turbine inlet temperature (T 14) and pressure (P 14), as well as the corresponding value for the mass flow rate (\(\dot{m}_{\text{wf}}\)). It also indicated the condition of the fluid at this state, which can be subcritical or supercritical.

Table 10 Optimized properties for non-recuperative ORC
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Table 11 Optimized properties for recuperative ORC
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The turbine inlet temperature and pressure vary significantly depending on the chosen working fluid. Some fluids, such as n-dodecane enter the turbine with an absolute pressure of 3.65 kPa, which is much lower than the atmospheric pressure. On the other hand, the highest pressure is required by the R717 (13.6 MPa). The mass flow rate also varies widely, from 1.03 kg/s for water to 13.24 kg/s for R502. Most substances with critical temperature lower than the waste heat source temperature reached supercritical condition. This is due to the better matching between the heat source temperature curve and working fluid temperature curve [14]. These results can be better observed in Fig. 7.

Fig. 7
figure 7

Optimization results

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As an overall result, n-butane presented the best results with a net power output of 409.3 kW and a cycle efficiency of 15.25%. Following this, R245fa comes with 408.5 kW of net power output and 15.22% of efficiency. Both reached supercritical condition. The poorest results were from R502 with a power output of 102.9 kW and 3.83% efficiency. Carnot efficiency is used to evaluate how far the ORC is from thermodynamic upper limit. The Carnot efficiency \(\eta_{\text{Carnot}}\) is defined in Eq. (23).

$$\eta_{\text{Carnot}} = 1 - \frac{{T_{\text{L}} }}{{T_{\text{H}} }},$$
(23)

T L is the condensation temperature, the lowest temperature of the cycle. T H is the temperature in which the heat is transferred to the cycle. The mean thermodynamic temperature of the exhaust gases is defined in Eq. (24).

$$T_{\text{mean}} = \frac{\Delta h}{\Delta s},$$
(24)

\(\Delta h\) is the enthalpy drop and \(\Delta s\) is the entropy variation of flow. T mean = 140.9 °C for the ORC optimized with n-butane as working fluid and T L = 50 °C (temperature at condenser). Considering T H = T mean, the corresponding Carnot efficiency is 21.95%. Therefore, the ORC with best results has an efficiency relative to Carnot (\(\eta_{\text{ORC}} /\eta_{\text{Carnot}} )\) of 69.48%, which is a notorious result.

2.5.6 Comparison between ORC-assisted and chiller-assisted power plants

To compare the results from both proposals, it was necessary to equalize the time periods. The use of an absorption chiller resulted in a gain of 1.81 MW for a period of 2669.41 s. On a daily basis, the result is equivalent to an average gain of 55.92 kW which represents an increase of 0.06% in the efficiency of the studied plant. The comparison results can be seen in Table 12.

Table 12 Comparison between ORC-assisted and chiller-assisted power plants
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3 Conclusion

The objective of this work was to compare two possible proposals to increase the power generated by an already existing conventional Rankine cycle-based power plant. The two systems tested were an absorption chiller to reduce the temperature of condenser cooling water and an organic Rankine cycle, both powered by the recoverable (above dew point) waste heat from the boiler exhaust gases. ORC optimizations were conducted using 23 different fluids searching for the optimal expander inlet temperature and pressure to obtain the maximum net power generation.

For the specific conditions evaluated, the absorption chiller increased the power output in 1.81 MW for a period of 44 min and 29 s, which represents a daily average gain of 55.92 kW. The Recuperative ORC with n-butane represented a gain of 409.3 kW in net power generation under the optimized conditions. These results indicate that, from a thermodynamic point of view, it is more advantageous to use an ORC coupled to the Rankine cycle-based power plant studied than an absorption chiller. Further studies need to be conducted to compare the economical benefits of both systems. Other technologies may also be applied to use the energy from the exhaust gases. Boiler design improvements such as addition of extra economizer or air pre-heaters are interesting options. These improvements, however, will change the temperature profile of the boiler, therefore, the evaluation of operational feasibility and changes in control system of this equipment will be required.