Heat recovery steam generators (HRSGs) are important components for industrial waste heat recovery, and any changes in their design directly affect the performance of the steam cycle, and thus the performance of the combined cycle power plant. The present research is focused on the design of a HRSG, including a dual-pressure steam generator cycle that is intended for use in a typical gas turbine unit having a power output of 60 MW. The results show that the maximum heat transfer occurs in the evaporator section for high-pressure levels, and in the economizer section for low-pressure levels. The optimum design pressure for the high-pressure level of the steam generator was found to be 100 bar, while that for the low-pressure level was found to be 10 bar. An exergy analysis was performed for the different processes and components of the steam cycle, including the HRSG. Calculations show that the overall exergy loss is about 35%, out of which 16% is lost in flue gasses and 10% is lost in heat exchanger, and the remainder was converted into 35 MW of useful work.

1 INTRODUCTION

1.1 Importance of waste heat

In many countries, a large portion of the total energy demand is consumed in industrial processes. For example, studies estimated that about 20–50% of the industrial energy consumption is discharged as heat. Waste heat recovery methods have shown a great development; however, there is still a big potential in their improvement [1]. Typically, there is greater potential to reuse high-temperature waste heat, where high is defined as >400°C, medium as 100–400°C and low as <100°C [2].

Combustion or burning processes are the most widespread thermal treatments applied to different types of waste, including municipal solid waste, solid refuse fuels, industrial waste (IW) and industrial hazardous waste [3]. The burning of waste is generally associated with energy recovery in the form of electricity and/or heat production. Industrial heat waste is frequently discarded by burning without energy recovery, because energy recovery for this type of waste is quite complex due to the presence of several pollutants in the generated flue gas. Even in such cases, energy recovery can be beneficial as it reduces the operating costs and lowers the consumption of external energy [4] and consequently should be applied whenever possible.

At present, waste thermal treatments are attractive options due to the possibility of recovering a significant amount of energy as a result of recent technological developments [4]. Pavlas [5] states that the thermal treatment of waste with energy recovery is a desirable source of renewable energy, and waste that can be used as a fuel is no longer classified as a problem. Recovering energy from waste heat primarily contributes to energy savings in conventional utility systems [6].

Natural fossil fuels account for 67% of global electricity generation. The major sources of energy and the percent share of total US electricity generation in 2015 were as follows: coal (33%), natural gas (33%), and nuclear (20%) [7]. These energy sources typically use boilers and/or steam turbine systems to generate steam for electricity generation. Most, but not all, of the heating systems in energy generation are used to service boilers that produce hot water or steam. All major industrial energy users consume a large portion of their fossil fuel in steam production: food processing (57%), chemicals (42%), pulp and paper (81%), petroleum refining (23%) and primary metals (10%) [8].

Waste heat recovery systems produce power by consuming the heat energy lost to the surroundings from thermal processes, with no additional fuel input. For marine vessels, about 50% of the total fuel energy supplied to the diesel power plants is lost to the surroundings [9]. Researchers investigated several different waste heat recovery systems to better understand the potential recovery efficiencies and their suitability to marine applications and concluded that the suitability of the waste heat recovery system changes depending on the temperature range of interest.

1.2 Design of waste HRSG

With population growth and rapid increases in industrial uses, the demand for electrical energy is increasing significantly in both developed and developing countries. Thus, reducing the costs of electricity and lowering pollutant emissions became the prime motivation for researchers and engineers to look for other more efficient and environmentally friendly technologies for power generation [9–11] . Among these technologies are waste heat recovery power plants consisting of a gas cycle, Brayton cycle and steam cycle, or Rankine cycle system linked by a HRSG. In these systems, the exhaust heat from the flue gases of the gas turbine is recovered in a HRSG to produce steam at a suitable pressure and temperature. Then, a steam turbine is used to produce electricity.

HRSGs are classified into single-, dual- and triple-pressure types, depending on the number of drums in the boiler. Dual-pressure HRSGs are widely used, as they offer higher efficiencies than single-pressure systems and a lower investment cost than triple-pressure HRSGs [12] and [13].

The thermal efficiency and power generation capacity of combined cycle power plants are dependent on the design of the HRSG. Thus, the HRSG must be carefully designed to take full advantage of the heat exchanged and to improve the overall performance of the plant [14].

Table 1 shows the temperature ranges and characteristics for IW heat sources [15].

The key challenge in designing a HRSG is to ensure the maximum utilization of the gas turbine exhaust heat (heat from the exhaust gases) in the heat recovery system to generate steam with the minimum heat exchange area to make the cycle more efficient and economical.

Many researchers focused on improving the performance of combined cycle power plants by properly designing the HRSG. P. K. Nag designed a HRSG for saturated steam for a combined gas and steam power cycle with minimum irreversibility [16] and concluded that the entropy is reduced when the HRSG is operated at full load. Franco studied the thermodynamic analysis to design the operating parameters for different configurations of HRSG systems to minimize the exergy losses considering only the irreversibility due to the difference in temperature between the cold and hot fluids [17]. The HRSG forms a major part of the steam system. In the combined cycle mode, the efficiency of the combined gas turbine-plus-HRSG system can reach 55–60% (lower heating value basis) with modern advanced machines, while in cogeneration mode, the system efficiency can be as high as 75–85%.

Recent trends in HRSG design include multiple-pressure units for maximum energy recovery, the use of high-temperature superheaters or reheaters in combined cycle plants and auxiliary firing for efficient steam generation.

The first step in designing a HRSG is to calculate the steam generation ability, and gas and steam temperature profiles. The flow rate and exit steam temperatures can be assumed based on the specifications for a conventional fired steam generator. The choice of these two factors then affects the sizes of the superheater, evaporator and economizer. Based on the sizes of evaporators that can be economically constructed and transported, the pinch and approach points for unfired HRSGs are usually in the range of 10–15°C. In the case where specifically less steam is desired, such as in a multiple-pressure HRSG generation system where there is more low-pressure steam then high-pressure steam, a larger pinch and approach may be used [18].

1.3 Exergy

Exergy follows from the Second Law of Thermodynamics and is a property that enables us to define the maximum useful work of a given amount of energy at a specified state. Exergy investigations of energy use were first introduced in the USA [19] and are now widely used in the design, simulation and performance evaluation of thermal and thermochemical systems.

From these analyses, it is known that the heat exchanger and combustor are the main parts contributing to the loss of energy, and the exergy efficiency is lower than the energy efficiency. Exergy is always evaluated with respect to a reference environment, which is in a stable state, acts as an infinite system, and is a sink for heat and materials. In this approach, the temperature T_o and pressure P remains constant, as recommended by [20]. In the present study, T₀ = 25°C and P₀ = 100 kPa [21].

2 FUNDAMENTALS OF DUAL-PRESSURE HRSG

Heat transfer is dependent on the mass flow of hot gases, mass flow of water or steam, temperature difference and surface area. A single-pressure HRSG can recover heat up to a specific level. If more heat is to be recovered, then the area of the heat exchanger increases, which soon becomes economically unfeasible. For this reason, designers often use dual-pressure HRSG systems to recover more heat and to reduce the stack temperature. In dual-pressure HRSG systems, two stages of pressure are used, where one is high and the other is relatively low. The flow rates of both stages may be different, depending on the requirements.

A dual-pressure HRSG is a series of heat exchangers arranged to maximize the amount of heat recovered and consists of three heat exchangers, namely, the economizer, evaporator and superheater, for both the high- and low-pressure levels. In the present research, three economizers were used for the high-pressure cycle, in such a way that reduces the size required, because by splitting the economizers and exposing them to the higher temperature of the exhaust gases results in smaller area for economizer. The arrangement of high- and low-pressure heat exchangers are such that the heat transferred to all sections is properly distributed to extract the maximum amount of heat from the flue gases. An economizer is used to heat the water close to the saturation temperature, evaporators are used to produce saturated steam, and superheaters are used to produce superheated steam. The heat exchangers are in the form of bundles of tubes placed in staggered arrangement to increase the heat transfer coefficient. The flow of the working fluid (water or steam) in the pipes is horizontal and the flow of exhaust gasses is vertical, which means each heat exchanger in the HRSG can be considered to be a cross flow heat exchanger as shown in Figure 1. The fluid in the evaporator undergoes forced circulation by pump for both pressure levels. Steam drums are used in HRSG to separate the water and steam from the evaporator. The arrangement of the evaporator is such that the wet steam collects in the steam drum by the natural law of mass transfer.

Figure 1.

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Schematic of the waste heat recovery cycle.

 

 

 

Figure 1 shows a schematic view of a gas turbine power plant with a typical dual-pressure HRSG and steam turbine system. Both open and closed feed water heat is used to preheat the water before it enters the heat exchanger. In the turbine after expansion, some bleeding is performed, which is used to heat the water in the open feed water heater.

The exhaust gases of the gas turbine have a low heat transfer coefficient as compared to the heat transfer coefficients of the working fluid in the evaporator and economizer; therefore, finned tubes are used in the HRSG. Annular fins are used to increase the heat transfer on the gas side. The most important parameters of HRSGs are the pinch point and approach point, which contribute to the effectiveness of the heat exchange. The pinch point is the difference between the gas temperature leaving the evaporator section and the temperature of the fluid entering the evaporator section. The approach point is the difference between the saturation and inlet temperatures of the fluid in the evaporator.

The selection of these two factors also affects the sizes of the superheater, evaporator and economizer. The pinch and approach points for unfired HRSGs are usually in the range of 10–15°C. If less steam is desired, then a larger pinch and approach may be used to simplify the heat transfer requirement in the heat exchanger by allowing a smaller surface area.

The behavior of a HRSG depends on the inlet and exit gas temperatures; thus, an arbitrary assumption regarding the exit gas temperature leads to a temperature cross situation. The right way to evaluate the design temperature profile is to assume pinch and approach points and perform the calculations. Variables that directly affect the steam generation and steam and gas temperature profiles are the approach and pinch points, as shown in Figure 2.

Figure 2.

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Approach and pinch points.

 

 

 

3 DESIGN APPROACH

3.1 Flow diagram

For design purposes, it is necessary to determine the magnitude (energy flow) and temperatures of the streams. The most reliable way to quantify the flows is through heat and mass balances. In simple cases, suitable heat and mass balances can be determined by hand calculation. In more complex cases, spreadsheet balances or specialist flow sheet simulation software may be the best way forward. Different calculation platforms have advantages and disadvantages, depending on the context.

Figure 3 represents the logical flow process used in the present design of the HRSG. In the figure, the flow is expanded into several steps. The first few steps are the decision-making steps for the HRSG. Decisions are taken regarding which method should be used under which condition, i.e. the selection of working fluid (water or Organic fluid), selection of design pressure, selection of single- or dual-pressure cycle. The single-pressure cycle recovers less heat as a result the exit temperature remains high, while in dual-pressure the heat recovered is more thus resulting in lower exit temperature. The next steps contain the design processes, i.e. the design of heat exchanger (economizer, evaporator and superheater), design of fins and design for preheating of the working fluid. The last step is the performance calculation of the cycle.

Figure 3.

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Logical flow diagram for the waste HRSG design.

 

 

 

3.2 Assumptions

The main assumptions used to develop the mathematical model are as follows:

• Both the water and exhaust gas sides of the system are assumed to be in steady state.

• The HRSG is unfired.

• There is no heat loss from the heat exchangers.

• To avoid the dew point, the stack temperature is assumed to be more than 100°C [22].

• The exhaust flue gases are forced to circulate by means of fans.

• The tubes are in a staggered arrangement with fins.

• The mass flow rate of the working fluid in the economizer, evaporator and superheater is assumed to be equal in each pressure level.

• The input flue gas condition flow rate and temperature are fixed.

• Pinch and approach point have to be assumed.

3.3 Mathematical modeling

When the actual processes can be represented by mathematical models, then the system can be simulated. The HRSG model depends on the mass flow rates, fluid dynamics, heat transfer and energy balance.

The following mathematical model was derived based on the assumptions above, and the main geometric variables are illustrated in Figure 4.

Figure 4.

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Geometry of the pipe.

 

 

 

3.4 Temperature profile

The first design step in the design process is to determine the temperature profile. The temperatures of the working fluid and exhaust gases at the inlet and outlet for each section can be calculated using the energy balance. The calculation should be done by using a suitable pinch point. Figure 2 shows the temperature distribution of heat exchanger.

3.5 Calculating the overall heat transfer coefficients

In practice, the heat exchanger tubes are made of high-conduction material with a small thickness. Therefore, it is reasonable to assume that ln (D₀/D_i)/2πkL ≅ 0

The present calculation shows that the Reynold number is always >3000, which means that the flow is turbulent. Thus, the Nusselt number (Nu) for turbulent flow in a smooth circular tube can be obtained from the Dittus–Boelter Equation [23].

The following is a correlation [24] for flow past a single circular cylinder: Figure 5 represents the arrangement of tubes in staggered tube arrangement.

Figure 5.

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Staggered tube arrangement.

 

 

 

In this section, the heat transfer coefficients have been calculated and the overall heat transfer coefficient is known

3.6 Area calculation

3.7 Feed water heater

A feed water heater is used to preheat the water for steam generation. There are two types of feed water heaters: open feed water heaters and closed feed water heaters. In an open feed water heater, the hot working fluid is directly mixed with the cold fluid before being transferred to the next step, but it is a key that both fluids are at the same pressure. In a closed feed water heater, a heat exchanger is used to transfer the heat from the hot to the cold fluid.

3.8 Power generated by steam turbine

3.9 Pump work and pressure losses in pipe

3.10 Overall plant efficiency

3.11 Effectiveness of heat exchanger

4 CASE STUDY

The design of the HRSG depends on the thermodynamic parameters of the exhaust gases from the gas turbine in simple cycle mode, such as the temperature, pressure and mass flow rate of gases. In this section, the above design method is applied to a case study based on the real data from a local electrical utility gas turbine power plant. The basic parameters are listed in Table 2.

By using the above data from an electric utility company a dual-pressure HRSG has been designed. Figure 6 shows the temperature profile for the steam and exhaust gases cycle, by having a pinch point of 10°C.

Figure 6.

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Temperature profile for dual-pressure.

 

 

 

5 EXERGY ANALYSIS

This section describes the method used to estimate the energy and exergy used, and the energy and exergy efficiencies for the heat exchanger and other items.

In the system, the working fluid water is circulated by the pump, enters the evaporator as compressed liquid at 105 bar and leaves as saturated vapor at the evaporator pressure. The dead-state condition is 1 bar of pressure at a temperature of 40°C.

Figure 7 shows the exergy distribution of each component of the cycle. The maximum available exergy from the flue gases is divided into two parts. One part is utilized in the heat exchanger while the other is lost. The heat exchanger passes the exergy to the turbine, which converts it into useful work heating the steam in the open feed water heater. The outlet of the turbine is collected in the condenser, which first closes the feed water heater to gain exergy and then opens the feed water heater. Finally, the fluid is pumped to the desired design pressure by adding exergy externally from pumps. Then, the fluid again circulates through the heat exchanger and the cycle continues.

Figure 7.

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Exergy distribution of the HRSG.

 

 

 

Figure 8 represents the exergy losses in the HRSG as a percentage of the exergy input. The exergy input is the exergy available from the flue gases. A large amount of exergy loss take place in the flue gases. The second largest exergy loss is in the heat exchanger due to the evaporator where the phase change happens.

Figure 8.

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Percentage of exergy losses in each component.

 

 

 

6 RESULTS AND DISCUSSION

Figure 9 shows the effect of design pressure vs. that of the heat exchanger (WHRSG) for a constant area. The x-axis shows possible design pressures from 50 to 100 bar. It can be seen that an increase in the design pressure leads to a decrease in the effectiveness of the heat exchanger because more heat is required to raise the steam to the same temperature. Thus, for the same area, the heat exchanger is less effective, which means that it cannot recover the required heat. Preheating is also a factor in the effectiveness. If the preheating is increased, then the effectiveness also increases because the inlet and outlet temperature difference is smaller.

Figure 9.

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Effect of the design pressure on the effectiveness of the HX.

 

 

 

Figure 10 shows the effect of the design pressure on the stack temperature of the heat exchanger (WHRSG) when the mass flow rate and area of the heat exchanger is kept constant. The x-axis shows the possible design pressure variation from 60 to 100 bar. It can be seen that an increase in the design pressure leads to an increase in the stack temperature because the heat transferred from the flue gases decreases, which causes less heat transfer between the fluids and an increase in the stack temperature. The preheating also has an impact on the stack exit temperature because, at a constant mass flow rate of working fluid and a constant area, the heat transfer becomes less between the working fluid and flue gas, which causes the stack exit temperature to increase.

Figure 10.

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Effect of the design pressure on the stack temperature.

 

 

 

Figure 11 shows the effect of the overall plant efficiency and pinch point temperature for single-pressure and multipressure HRSGs. The x-axis shows the possible variation of pinch points from 20°C to 40°C. The overall plant efficiency decreases with increasing evaporator pinch point temperature difference. For single pressure, this decrement is ~0.64%. For a double-pressure HRSG, it is ~0.25%. This is because increasing the pinch point causes the steam production to decrease, which leads to a decrease in the overall plant efficiency. In the current analysis, the overall plant efficiency increased up to 52.4%.

Figure 11.

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Effect of the overall plant efficiency and pinch point.

 

 

 

Figure 12 shows the total heat transfer in the HRSG and the temperature of the flue gas. By increasing the input flue gas temperature, the heat exchange increases because, at higher temperatures, the enthalpy of the flue gas increases, which causes more heat transfer to take place.

Figure 12.

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Effect of the input flue gas temperature on the total heat exchanged in the HRSG.

 

 

 

Figure 13 shows the heat transfer rate in each heat exchanger. In the figure below, it is clear that the maximum heat transfer occurs in the evaporator for both pressure cycles due to higher latent heat of vaporization of water.

Figure 13.

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Heat transfer rate in each heat exchanger for both pressure levels.

 

 

 

7 CONCLUSION

This paper presents a complete design methodology for a waste HRSG. A typical case study for an Electric Utility Plant is also presented. By using a HRSG for recovering the waste heat from the exhaust of a 60-MW gas turbine, an additional power of 35.14 MW can be generated, thus increasing the overall plant efficiency from 33% to 52%. The heat transfer coefficients of gas are weak; therefore, the designed HRSG has a large area to produce steam at high pressure and temperature. Multiple-pressure steam generation should be considered to optimize energy recovery, particularly if high-pressure steam is generated. It is more efficient to use a dual-pressure cycle instead of a single-pressure cycle. The difference in overall plant efficiency between single- and dual-pressure at a constant pinch point, i.e. 20°C, was found to be 1.5%. A dual-pressure steam generator affects more than only the efficiency of the overall plant; it also ensures small variation at pinch points. The variation is about 1% for a pinch point range of 20–40°C. Pinch and approach points should be selected according to the need of the steam generated.

Exergy is an effective method that uses the conservation of mass and conservation of energy principles together with the second law of thermodynamics to design and analyze energy systems. The overall exergy loss is about 35%, out of which 16% is lost in flue gasses and 10% is lost in heat exchanger. This is due to the phase change in the evaporator section and isentropic efficiency of turbine.

ACKNOWLEDGEMENTS

The authors gratefully acknowledge Qassim University and SABIC, represented by the Deanship of Scientific Research, on the material support for this research under grant number 2857 during the academic year 1435 AH. 2014 AD.

NOMENCLATURE

• Q

  heat rate (kW)

• h

  enthalpy (kJ/kg)

• m

  mass flow rate (kg/s)

• U

  overall heat transfer coefficient (kW/m2K)

• A

  area (m2)

• D

  diameter (m)

• k

  thermal conductivity of pipe (W/mK)

• Nu_D

  Nusselt number

• Pr

  Prandtl number

• μ

  dynamic viscosity (N.s/m2)

• ρ

  density (kg/m3)

• V

  velocity (m/s)

• λ

  parametric length of tube (m)

• S_Q

  transverse tube spacing (m)

• S_L

  longitudinal tube spacing (m)

• r

  radius (m)

• N

  number of tubes

• n

  number of fins on the unit pipe

• HX

  Heat Exchanger

• NTU

  number of transfer units

• f

  Darcy coefficient of friction

• L

  length (m)

• W

  work done (kJ)

• Re

  Reynold number

• E

  Exergy (kJ)

• μbμb

  base fluid viscosity(N.sm2).base fluid viscosity(N.sm2).

• μwμw

  fluid viscosity at the wall of the pipe(N.sm2).fluid viscosity at the wall of the pipe(N.sm2).

Subscripts

• g

  gas

• i

  inlet

• l

  liquid phase

• o

  outlet

• cold(c)

  fluid at the cold side

• hot(h)

  fluid at the hot side

• b

  bleeding

• p

  pump

• t

  turbine

• s1

  high pressure cycle steam

• s2

  low pressure cycle steam

REFERENCES