**Gas Turbines in Unconventional Applications**

Jarosław Milewski, Krzysztof Badyda and Andrzej Miller *Institute of Heat Engineering at Warsaw University of Technology Poland*

### **1. Introduction**

120 Efficiency, Performance and Robustness of Gas Turbines

Rosen MA, Dincer, I, 2003 a. Thermoeconomic analysis of power plants: an application to a

Stewart W and Patrick A. 2000. Air temperature depression and potential icing at the inlet of stationary combustion turbines. *ASHRAE Transactions* 106 (part 2):318-327. Szargut, J., Morris, D.R., and Steward, F.R. (1988) 'Exergy analysis of thermal, chemical and

Tillman TC, Blacklund DW, Penton JD. 2005. Analyzing the potential for condensate carry-

Tyagi S.K., Chen G.M., Wang Q. and Kaushik, S.C., (2006) 'A new thermoeconomic

Zaki GM, Jassim, RK, Alhazmy, MM, 2007 "Brayton Refrigeration Cycle for gas turbine

refrigeration cycle' Int. J. Refrigeration, Vol. 29, No. 7, pp 1167-1174. Zadpoor AA, Golshan AH, 2006. Performance improvement of a gas turbine cycle using a

desiccant-based evaporative cooling system. Energy 31: 2652-2664.

No. 17, pp 2743-2761.

DE-05-6-3:555-563.

1306.

metallurgical process'. Hemisphere, NY, USA.

coal fired electrical generating station. *Energy Conversion and Management*, Vol. 44,

over from a gas cooling turbine inlet cooling coil. *ASHRAE transactions* 111 (part 2)

approach and parametric study of an irreversible regenerative Brayton

inlet air cooling, International Journal of Energy Research, Vol. 31, No. 13 pp 1292-

The Chapter presents unconventional gas turbine applications. Firstly some selected non-MAST (Mixed Air and Steam Turbine) solutions are discussed – these are intended for smaller gas turbine systems, where the regenerative heat exchanger supplies energy for an additional thermal cycle, where it is utilised.

The gas turbine cycle (Brayton) may be coupled with several other thermal engines (like another Brayton, Diesel, Kalina, and Stirling). Those hybrid systems have several previously unrecognised advantages. They may find applications in some market niches.

The next section describes hydrogen-fuelled gas turbine solution. Big international programme – WE-NET – is discussed. Several hydrogen-fuelled gas turbine concepts based on those programmes are proposed: Westinghouse, Toshiba, Graz, New Rankine. The section provides description of them all, including specification of possible efficiency values. The development programmes themselves are also reviewed. This part of the text describes also potential combination of a hydrogen-fuelled gas turbine and a nuclear power generation unit which might be used to cover peak load power demands in a power system.

Last section of the chapter discusses integration of a fuel cell into a gas turbine system. High temperature fuel cells can play a role similar to a combustion chamber but simultaneously generating additional power. Fuel cell hybrid systems for both high-temperature types of fuel cells – Solid Oxide Fuel Cell (SOFC) and Molten Carbonate Fuel Cell (MCFC) – are proposed. Additionally, some specific properties of the MCFC can be used to reduce carbon dioxide emissions from the gas turbine itself.

Gas turbine systems, particularly combined cycle units, are among the most popular power systems in the modern world. This results from the very fast technical progress allowing to gradually increase the parameters at the turbine inlet as well as unit outputs. There is also a parallel development trend of searching for new unconventional solutions, which would allow to achieve efficiencies higher than enabled by a simple cycle. Scheme of the simple cycle process is presented in Fig. 1. Simple cycle efficiency in many cases is too low.

Internal power *Ni* of a gas turbine can be obtained by an analytical approach by using the relation 1, which is obtained with assumption that the process is real (contained losses) and working fluid is modelled as semi-ideal gas.

T

G

(4)

4

V

CC

Gas Turbines in Unconventional Applications 123

2 3

Fig. 2. Simple cycle gas turbine diagram: C – compressor, T – turbine, G – electric generator,

Fig. 3. Specific power output and thermal efficiency of a simple cycle gas turbine unit as

*<sup>η</sup><sup>c</sup>* <sup>=</sup> *Ni Np*

Characteristics of a simple cycle gas turbine unit based on the relationships presented above are shown below. Performance calculations were carried out at variable pressure ratio Π*<sup>K</sup>* and temperature *T*<sup>3</sup> (at turbine inlet, without taking into consideration cooling losses, i.e. Δ*G* = 0). Constant values of polytropic efficiencies of turbine *η<sup>T</sup>* = 0.88 and compressor *η<sup>K</sup>* = 0.88, pressure losses factor *ε* = 0.95 were used. Ambient conditions (pressure and temperature)

The first curve (Fig. 3) illustrates key performance parameters of a simple cycle gas turbine unit (specific power output and efficiency) as functions of pressure ratio and Turbine Inlet Temperature (TIT). It can be seen that in a real system maximal specific power output is

Another curve shows relation between efficiency and a specific power output (Fig. 4). This curve was obtained with the same relationships and assumptions as for Fig. 3. The range of investigated temperatures was extended (while keeping the assumption that there is no turbine cooling, which results with performance figures being somewhat exaggerated). Points on individual curves denote selected pressure ratio values. Analysis of the cooling system

It is possible to considerably increase efficiency and other performance figures of a gas turbine unit at the expense of introducing new components and making the flow system more complex. Therefore, except for simple gas turbines, also complex units with more

impact on performance of a gas turbine can be found for example in Nat (2006).

functions of the pressure ratio. Dashed lines denote specific power output.

and thermal efficiency is a ratio of internal power and fuel power input:

were assumed according the ISO conditions:*T*<sup>0</sup> = 288 K (15°C).

achieved at lower pressure ratio then the maximal efficiency.

C

CC – combustion chamber, P – fuel pump, V – bypass valve

P

1 0 (surroundings)

Fig. 1. Thermal process of a simple cycle gas turbine

$$\begin{split} N\_{l} &= N\_{T} - N\_{\mathbf{K}} \\ &= \mathbf{G}\_{T} \cdot \mathbf{c}\_{p,T} \cdot T\_{3} \left( 1 - \mathbf{x}\_{T} \right) \cdot \eta\_{T} - \mathbf{G}\_{\mathbf{K}} \cdot \mathbf{c}\_{p,K} \cdot T\_{1} \cdot \left( \mathbf{x}\_{\mathbf{K}} - 1 \right) \cdot \frac{1}{\eta\_{\mathbf{K}}} s \\ &= \mathbf{G}\_{\mathbf{K}} \cdot \mathbf{c}\_{p,K} \cdot T\_{1} \cdot \left[ \frac{1}{1 - \beta} \cdot \overline{\mathbf{c}}\_{p} \cdot \Theta \cdot \left( 1 - \mathbf{x}\_{T} \right) \cdot \eta\_{T} - \left( \mathbf{x}\_{\mathbf{K}} - 1 \right) \cdot \frac{1}{\eta\_{\mathbf{K}}} \right] \end{split} \tag{1}$$

where: *NT*, *GT*, *NK*, *GK* stand for internal power and fluid inlet mass flow in turbine and compressor, respectively, *η<sup>T</sup>* – turbine internal efficiency, *η<sup>K</sup>* – compressor polytropic efficiency, *cp*,*T*, *cp*,*<sup>K</sup>* – averaged isobaric specific heat capacity for the working fluid in turbine and compressor, *kK*, *kT* – averaged heat capacity ratios (isentropic exponents) for compressor and turbine for the flue gas and air. Other symbols used in Eq. 1 are:

$$\begin{aligned} \tilde{\varepsilon}\_{p} &= \frac{c\_{p,T}}{c\_{p,K}}; \quad \Theta = \frac{T\_3}{T\_1} \\ \mathbf{x}\_T &= \frac{1}{\Pi\_T^{m\_T}}; \, m\_T = \frac{k\_T - 1}{k\_T}; \, \Pi\_T = \frac{p\_3}{p\_4} \\ \mathbf{x}\_K &= \Pi\_K^{m\_K}; \, m\_K = \frac{k\_T - 1}{k\_K}; \, \Pi\_K = \frac{p\_2}{p\_1} = \Pi \\ \Pi &= \Pi\_K = \frac{p\_2}{p\_1}; \quad \Pi\_T = \frac{p\_3}{p\_4} \quad \Pi\_T = \varepsilon \Pi\_K \\ \beta &= \frac{G\_p - \Delta G}{G\_T} \end{aligned} \tag{2}$$

where: *Gp* – fuel mass flow delivered to a combustion chamber, Δ*G* – air mass flow delivered for cooling purposes of the hottest parts of gas turbine and bearings, and leakages losses, Π*T*, Π*<sup>K</sup>* are pressure ratios of outlet and inlet of compressor and turbine, and losses factor *ε* describes combination of pressure losses at compressor inlet, combustion chamber and turbine outlet.

Subscripts of working fluid parameters (pressure and temperature) indicate location in reference to the turbine diagram (Fig. 2). Specific internal power by definition is expressed as:

$$N\_{\circ} = \frac{N\_{\text{i}}}{G\_{\text{K}}} \tag{3}$$

2 Will-be-set-by-IN-TECH

p =p 2 2 max

T =T 3 3 max 3id

p0

p4

4s

4id

3

Äp<sup>2</sup>

4

Äp4

p1

Äp<sup>0</sup>

<sup>=</sup> *GT* · *cp*,*<sup>T</sup>* · *<sup>T</sup>*<sup>3</sup> (<sup>1</sup> <sup>−</sup> *xT*) · *<sup>η</sup><sup>T</sup>* <sup>−</sup> *GK* · *cp*,*<sup>K</sup>* · *<sup>T</sup>*<sup>1</sup> · (*xK* <sup>−</sup> <sup>1</sup>) · <sup>1</sup>

where: *NT*, *GT*, *NK*, *GK* stand for internal power and fluid inlet mass flow in turbine and compressor, respectively, *η<sup>T</sup>* – turbine internal efficiency, *η<sup>K</sup>* – compressor polytropic efficiency, *cp*,*T*, *cp*,*<sup>K</sup>* – averaged isobaric specific heat capacity for the working fluid in turbine and compressor, *kK*, *kT* – averaged heat capacity ratios (isentropic exponents) for compressor

> ; Θ = *<sup>T</sup>*<sup>3</sup> *T*1

; *mT* = *kT*−<sup>1</sup>

*<sup>K</sup>* ; *mK* <sup>=</sup> *kK*−<sup>1</sup>

; <sup>Π</sup>*<sup>T</sup>* <sup>=</sup> *<sup>p</sup>*<sup>3</sup>

where: *Gp* – fuel mass flow delivered to a combustion chamber, Δ*G* – air mass flow delivered for cooling purposes of the hottest parts of gas turbine and bearings, and leakages losses, Π*T*, Π*<sup>K</sup>* are pressure ratios of outlet and inlet of compressor and turbine, and losses factor *ε* describes combination of pressure losses at compressor inlet, combustion chamber and turbine

Subscripts of working fluid parameters (pressure and temperature) indicate location in reference to the turbine diagram (Fig. 2). Specific internal power by definition is expressed

> *Nj* <sup>=</sup> *Ni GK*

2

0 1

2s

S

<sup>1</sup> <sup>−</sup> *<sup>β</sup>* · *<sup>c</sup>*¯*<sup>p</sup>* · <sup>Θ</sup> · (<sup>1</sup> <sup>−</sup> *xT*) · *<sup>η</sup><sup>T</sup>* <sup>−</sup> (*xK* <sup>−</sup> <sup>1</sup>) · <sup>1</sup>

*kT* ; <sup>Π</sup>*<sup>T</sup>* <sup>=</sup> *<sup>p</sup>*<sup>3</sup>

*kK* ; <sup>Π</sup>*<sup>K</sup>* <sup>=</sup> *<sup>p</sup>*<sup>2</sup>

*<sup>p</sup>*<sup>4</sup> Π*<sup>T</sup>* = *ε*Π*<sup>K</sup>*

*p*4

*p*1

*ηK*

*ηK* 

= Π (2)

(3)

*s* (1)

T

T3

T1

1

and turbine for the flue gas and air. Other symbols used in Eq. 1 are:

*<sup>c</sup>*¯*<sup>p</sup>* <sup>=</sup> *cp*,*<sup>T</sup> cp*,*<sup>K</sup>*

*xT* <sup>=</sup> <sup>1</sup> Π*mT T*

*xK* = <sup>Π</sup>*mK*

*p*1

<sup>Π</sup> <sup>=</sup> <sup>Π</sup>*<sup>K</sup>* <sup>=</sup> *<sup>p</sup>*<sup>2</sup>

*<sup>β</sup>* <sup>=</sup> *Gp* <sup>−</sup> <sup>Δ</sup>*<sup>G</sup> GT*

Fig. 1. Thermal process of a simple cycle gas turbine

= *GK* · *cp*,*<sup>K</sup>* · *T*<sup>1</sup> ·

*Ni* = *NT* − *NK*

outlet.

as:

T2

T4

Fig. 2. Simple cycle gas turbine diagram: C – compressor, T – turbine, G – electric generator, CC – combustion chamber, P – fuel pump, V – bypass valve

Fig. 3. Specific power output and thermal efficiency of a simple cycle gas turbine unit as functions of the pressure ratio. Dashed lines denote specific power output.

and thermal efficiency is a ratio of internal power and fuel power input:

$$
\eta\_{\mathcal{C}} = \frac{N\_{\mathcal{I}}}{N\_p} \tag{4}
$$

Characteristics of a simple cycle gas turbine unit based on the relationships presented above are shown below. Performance calculations were carried out at variable pressure ratio Π*<sup>K</sup>* and temperature *T*<sup>3</sup> (at turbine inlet, without taking into consideration cooling losses, i.e. Δ*G* = 0). Constant values of polytropic efficiencies of turbine *η<sup>T</sup>* = 0.88 and compressor *η<sup>K</sup>* = 0.88, pressure losses factor *ε* = 0.95 were used. Ambient conditions (pressure and temperature) were assumed according the ISO conditions:*T*<sup>0</sup> = 288 K (15°C).

The first curve (Fig. 3) illustrates key performance parameters of a simple cycle gas turbine unit (specific power output and efficiency) as functions of pressure ratio and Turbine Inlet Temperature (TIT). It can be seen that in a real system maximal specific power output is achieved at lower pressure ratio then the maximal efficiency.

Another curve shows relation between efficiency and a specific power output (Fig. 4). This curve was obtained with the same relationships and assumptions as for Fig. 3. The range of investigated temperatures was extended (while keeping the assumption that there is no turbine cooling, which results with performance figures being somewhat exaggerated). Points on individual curves denote selected pressure ratio values. Analysis of the cooling system impact on performance of a gas turbine can be found for example in Nat (2006).

It is possible to considerably increase efficiency and other performance figures of a gas turbine unit at the expense of introducing new components and making the flow system more complex. Therefore, except for simple gas turbines, also complex units with more

T

flue gas

G

G

CC

Gas Turbines in Unconventional Applications 125

C T

air air

Fig. 5. Diagram of a combined cycle composed of a gas turbine and an air turbine supplied

a simple cycle solution. When compared to a classic GTCC, the Brayton-Brayton techenology requires less auxiliary equipment. It also needs less space and has lower investment cost. Systems of this type are not extensively analysed in the literature. Discussion presented in this chapter is based on the analyses presented in detail in Baader (n.d.). Performance of each gas turbine utilised in the system may be determined according to the rules and relations

Performance of the entire system depends on parameters (selection criteria) of the heat exchanger and air-based turbine unit. Calculations carried out in order to determine

1. The air cycle is designed to maximally utilise energy carried with the flue gas exhausted from the gas turbine unit. The flue gas is cooled as far as possible, but to temperatures no lower than 200◦C. Efficiency of the flue/air heat exchanger is 80% and the pinch point

2. Pressure ratio of the air turbine is designed to enable achieving highest possible internal

3. Assumed compressor and turbine polytropic efficiencies are 88% (just like for the simple

4. Assumed pressure losses at the heat exchanger for both air and flue gas is 3.4%, which

5. Just like in case of previous calculations no impact of internal gas turbine cooling systems

Other assumptions and the relationships used followed the rules previously presented for the simple cycle. Results are shown in the charts – characteristic curves for the Brayton-Brayton system. Fig. 7 shows relation between the system's efficiency and specific power output. Values shown here indicate that a considerable increase in reference to a simple cycle system

Maximum specific power occurs at a lower pressure ratio value than for a simple cycle system, but at higher values than for a system with regeneration. Specific power is much higher than

performance of a system shown in Fig. 5 were based on the following assumptions:

C

with heat recovered from the flue gas flow. (Brayton-Brayton)

results with the pressure loss coefficient of 0.928.

on system's performance is taken into account.

in case of a simple cycle or a regenerative system.

may be expected (compare to Fig. 2).

presented in the introduction.

temperature is 30◦C.

power.

cycle).

fuel

air

Fig. 4. Thermal efficiency of a simple cycle gas turbine unit as a function of specific power and turbine inlet temperature.

sophisticated working cycle are developed and build. These may involve for example heat regeneration, inter-stage cooling and re-heating. Such solutions require introduction of new components: regenerative (or recuperative) heat exchanger, gas coolers, combustion chamber cooling systems or inter-stage heaters. Also the compressor and turbine need to be split into high- and low-pressure parts (or even more sections).

In parallel to development of simple cycle gas turbines, also combined cycle solutions are developed, among them units with additional thermal circuit. Technical progress in gas turbine development is coupled with development of combined cycle systems, including solutions with combination of different thermal engine types. Combination of a gas turbine thermal cycle with an intermediate- or low-temeprature steam power plant cycle may be seen as the most efficient method to increase the efficiency of a process involving gas turbines used in industrial practice. Nowadays gas turbine combined cycle solutions are among the most intensely developed power generation technologies. "Conventional" Gas Turbine Combined Cycle or GTCC solutions may reach a nominal efficiency reaching 60%.

It is however not possible to achieve such high efficiencies at intermediate output GTCC units. It may prove more feasible to utilise simpler systems at lower scales. Specific investment cost for low-output GTCC unit is quite high, while the efficiency may be considerably lower. They do display high reliability though. One method to improve performance of low-output GTCCs may be utilising recuperation, inter-stage compressor cooling or working agent integration, e.g. utilising air-steam mixture as a working agent.

This last concept utilised in gas turbine systems has been proposed in various forms and under different names. One of the used names is Mixed Air and Steam Turbine (MAST). Another name is Cheng Cycle, and General Electric utilises name Steam Injected Gas Turbine – STIG. The gas turbine concepts with air humidification are known as Humid Air Turbines (HAT), Cascaded Humidified Advance Turbines (CHAT) or Wet Compression systems.

## **2. Gas turbines with other thermal engines**

### **2.1 Brayton-Brayton**

The Brayton-Brayton cycle (diagram shown in Fig. 5) is a combination of two simple-cycle gas turbines. Working agent in one of them is flue gas, and in the other – air. The turbines are interconnected with a high-temeprature air/flue gas heat exchanger. As the air turbine utilises exhaust heat, this solution enables to considerably improve total efficiency when compared to 4 Will-be-set-by-IN-TECH

Fig. 4. Thermal efficiency of a simple cycle gas turbine unit as a function of specific power

sophisticated working cycle are developed and build. These may involve for example heat regeneration, inter-stage cooling and re-heating. Such solutions require introduction of new components: regenerative (or recuperative) heat exchanger, gas coolers, combustion chamber cooling systems or inter-stage heaters. Also the compressor and turbine need to be split into

In parallel to development of simple cycle gas turbines, also combined cycle solutions are developed, among them units with additional thermal circuit. Technical progress in gas turbine development is coupled with development of combined cycle systems, including solutions with combination of different thermal engine types. Combination of a gas turbine thermal cycle with an intermediate- or low-temeprature steam power plant cycle may be seen as the most efficient method to increase the efficiency of a process involving gas turbines used in industrial practice. Nowadays gas turbine combined cycle solutions are among the most intensely developed power generation technologies. "Conventional" Gas Turbine Combined

It is however not possible to achieve such high efficiencies at intermediate output GTCC units. It may prove more feasible to utilise simpler systems at lower scales. Specific investment cost for low-output GTCC unit is quite high, while the efficiency may be considerably lower. They do display high reliability though. One method to improve performance of low-output GTCCs may be utilising recuperation, inter-stage compressor cooling or working

This last concept utilised in gas turbine systems has been proposed in various forms and under different names. One of the used names is Mixed Air and Steam Turbine (MAST). Another name is Cheng Cycle, and General Electric utilises name Steam Injected Gas Turbine – STIG. The gas turbine concepts with air humidification are known as Humid Air Turbines (HAT), Cascaded Humidified Advance Turbines (CHAT) or Wet Compression systems.

The Brayton-Brayton cycle (diagram shown in Fig. 5) is a combination of two simple-cycle gas turbines. Working agent in one of them is flue gas, and in the other – air. The turbines are interconnected with a high-temeprature air/flue gas heat exchanger. As the air turbine utilises exhaust heat, this solution enables to considerably improve total efficiency when compared to

and turbine inlet temperature.

high- and low-pressure parts (or even more sections).

Cycle or GTCC solutions may reach a nominal efficiency reaching 60%.

agent integration, e.g. utilising air-steam mixture as a working agent.

**2. Gas turbines with other thermal engines**

**2.1 Brayton-Brayton**

Fig. 5. Diagram of a combined cycle composed of a gas turbine and an air turbine supplied with heat recovered from the flue gas flow. (Brayton-Brayton)

a simple cycle solution. When compared to a classic GTCC, the Brayton-Brayton techenology requires less auxiliary equipment. It also needs less space and has lower investment cost. Systems of this type are not extensively analysed in the literature. Discussion presented in this chapter is based on the analyses presented in detail in Baader (n.d.). Performance of each gas turbine utilised in the system may be determined according to the rules and relations presented in the introduction.

Performance of the entire system depends on parameters (selection criteria) of the heat exchanger and air-based turbine unit. Calculations carried out in order to determine performance of a system shown in Fig. 5 were based on the following assumptions:


Other assumptions and the relationships used followed the rules previously presented for the simple cycle. Results are shown in the charts – characteristic curves for the Brayton-Brayton system. Fig. 7 shows relation between the system's efficiency and specific power output. Values shown here indicate that a considerable increase in reference to a simple cycle system may be expected (compare to Fig. 2).

Maximum specific power occurs at a lower pressure ratio value than for a simple cycle system, but at higher values than for a system with regeneration. Specific power is much higher than in case of a simple cycle or a regenerative system.

Fig. 9. Ratio between the gas cycle pressure ratio and air cycle pressure ratio in a

CC

Gas Turbines in Unconventional Applications 127

2 3

exchanger. In fact such a restriction would result from the material properties – this effect would practically rule out systems with low gas cycle pressure ratio. Practically only systems where the gas cycle parameters are similar to those used in simple cycle systems, i.e. those close to the parameters range enabling maximum specific power outputs and efficiencies (see

One of the interesting proposals presented in the literature is the so-called Brayton-Diesel

The Brayton-Diesel system (Fig. 10) is a combination of a simple cycle gas turbine with a heat exchanger and a reciprocating expander. Working agents are exhaust gas and air. Part of the air flow from the compressor is supplied to the combustion chamber, while some air is transferred into a heat exchanger where it is heated by the gas turbine exhaust. Then this flow expands in a reciprocating expander and is further fed into the low-pressure stages of the gas turbine, mixing with the exhaust gases. It was assumed that the mixing process is isobaric.

Application of the Brayton-Diesel system allows to increase maximum power output of a single unit by some 6–12% and increase maximum efficiency by 1.3–3.6% in the investigated

T

flue gas

G

air

G

4

Brayton-Brayton system as a function of the gas cycle pressure ratio.

5

flue gas

The resulting mixture further expands to the turbine exhaust pressure.

C

fuel

air

Fig. 3 and Fig. 4), should be seen as reasonable.

Fig. 10. Brayton-Diesel system flowchart

**2.2 Brayton-Diesel**

system Poullikkas (2005).

temperature range (900–1,300◦C).

1

0

Fig. 6. Thermal efficiency of a Brayton-Brayton system as a function of specific power output and Turbine Inlet Temperature. Numerical values shown in the chart denote the gas turbine pressure ratio.

Fig. 7. Thermal efficiency of a Brayton-Brayton system as a function of gas cycle's pressure ratio and turbine inlet temperature

Fig. 8. Share of the internal power generated in the air turbine in a total internal power of a Brayton-Brayton system as a function of gas cycle's pressure ratio and turbine inlet temperature.

Fig. 7 shows relation between the efficiency of a Brayton-Brayton system and the pressure ratio in the gas cycle. The next figure (Fig. 8) illustrates distribution of internal power between the gas and air cycles. High share of the air cycle at low gas cycle pressure ratio values results from the high temperature of exhaust gas delivered to the heat exchanger downstream from the gas turbine unit in such a case. This allows to achieve high pressure ratio in the air cycle – as already mentioned this value is optimised to achieve highest possible output. As shown in Fig. 9 the pressure ratio of the air cycle at low pressure ratio of gas cycle is high ΠKP/Π much lower than one). The results in this area should be seen as exaggerated, as the model does not include any limit of maximum (achievable) air temperature downstream from the heat 6 Will-be-set-by-IN-TECH

Fig. 6. Thermal efficiency of a Brayton-Brayton system as a function of specific power output and Turbine Inlet Temperature. Numerical values shown in the chart denote the gas turbine

Fig. 7. Thermal efficiency of a Brayton-Brayton system as a function of gas cycle's pressure

Fig. 8. Share of the internal power generated in the air turbine in a total internal power of a

Fig. 7 shows relation between the efficiency of a Brayton-Brayton system and the pressure ratio in the gas cycle. The next figure (Fig. 8) illustrates distribution of internal power between the gas and air cycles. High share of the air cycle at low gas cycle pressure ratio values results from the high temperature of exhaust gas delivered to the heat exchanger downstream from the gas turbine unit in such a case. This allows to achieve high pressure ratio in the air cycle – as already mentioned this value is optimised to achieve highest possible output. As shown in Fig. 9 the pressure ratio of the air cycle at low pressure ratio of gas cycle is high ΠKP/Π much lower than one). The results in this area should be seen as exaggerated, as the model does not include any limit of maximum (achievable) air temperature downstream from the heat

Brayton-Brayton system as a function of gas cycle's pressure ratio and turbine inlet

pressure ratio.

temperature.

ratio and turbine inlet temperature

Fig. 9. Ratio between the gas cycle pressure ratio and air cycle pressure ratio in a Brayton-Brayton system as a function of the gas cycle pressure ratio.

Fig. 10. Brayton-Diesel system flowchart

exchanger. In fact such a restriction would result from the material properties – this effect would practically rule out systems with low gas cycle pressure ratio. Practically only systems where the gas cycle parameters are similar to those used in simple cycle systems, i.e. those close to the parameters range enabling maximum specific power outputs and efficiencies (see Fig. 3 and Fig. 4), should be seen as reasonable.

### **2.2 Brayton-Diesel**

One of the interesting proposals presented in the literature is the so-called Brayton-Diesel system Poullikkas (2005).

The Brayton-Diesel system (Fig. 10) is a combination of a simple cycle gas turbine with a heat exchanger and a reciprocating expander. Working agents are exhaust gas and air. Part of the air flow from the compressor is supplied to the combustion chamber, while some air is transferred into a heat exchanger where it is heated by the gas turbine exhaust. Then this flow expands in a reciprocating expander and is further fed into the low-pressure stages of the gas turbine, mixing with the exhaust gases. It was assumed that the mixing process is isobaric. The resulting mixture further expands to the turbine exhaust pressure.

Application of the Brayton-Diesel system allows to increase maximum power output of a single unit by some 6–12% and increase maximum efficiency by 1.3–3.6% in the investigated temperature range (900–1,300◦C).

ST

LTHE HTHE

K

separator

4

HRSG

G

T

Fig. 13. Process diagram of a Brayton-Kalina system: C – compressor, T – turbine, G – generator, CC – combustion chamber, ST – steam turbine, K – Condenser, HRSG – heat recovery steam generator, HTHE – high temperature heat exchanger, LTHE – low

expander increases thermal efficiency and specific power of the system.

temperature in the system – the higher growth of those parameters.

the system and possibly low pressure ratio of the simple cycle.

1. Combination of a simple-cycle gas turbine with a heat exchanger and a reciprocating

2. Maximal thermal efficiency and maximal specific power increase; the higher the maximum

3. Maximal thermal efficiency grows slightly, the growth of the specific power is more

4. Optimum values of the simple cycle pressure ratio are lower for the combined cycle than

5. Using reciprocating expander is most profitable for possibly high maximal temperature of

6. Heat exchanger used in this system has to provide significant amount of heat to the air (Δ*Tpow* = 173. . . 470◦C) so it needs to have quite large heat exchange area which makes it

7. Discussed system can also be used in combined heat and power applications as the air-exhaust gas mixture flowing out of the heat exchanger has a quite temperature in range

While selecting design parameters of the Brayton-Diesel system also the economical analysis is very important. Increasing TIT and using very large heat exchanger could prove too

The Kalina system is a variant of the Organic Rankine Cycle (ORC). It utilises a mixture of water and ammonia as a working agent. Ratio between ammonia and water may be changed, depending on the process, which enables adjusting boiling and condensation temperatures. The Kalina cycle is based on the Rankine cycle with added distilling and absorption components (Fig. 14). Its main advantage is already mentioned ability to adjust boiling and

expensive when compared to the benefits resulting from the higher efficiency.

CC

2 3

C

fuel

air

temperature heat exchanger.

significant.

for the simple cycle.

expensive and bulky.

170...330◦C.

**2.3 Brayton-Kalina**

0 1

5

Gas Turbines in Unconventional Applications 129

flue gas

Fig. 11. Thermal efficiency of a combination of a gas turbine and reciprocating expander (configuration as in Fig. 10) as a function of specific power and TIT (T3). Numerical values shown in the chart denote gas turbine pressure ratios.

Fig. 12. Increase of a specific power output of a system with a reciprocating expander in comparison to the specific power output of a simple cycle gas turbine as a function of gas turbine pressure ratio and TIT (T3).

An interesting solution could be also supplying fuel to the reciprocating expander, which would also require combustion process there. This case is currently a subject of further investigation.

The Brayton-Diesel system can also be used for combined heat and power applications. Mixture of exhaust gas and air at the outlet of the heat exchanger has a temperature in range of 170–330◦C, so it can be used as a source of process heat.

Internal power of the combined cycle *Niz* was calculated according to the rules and relationships given in Miller (1984), by binding it to the parameters of the thermal process according to equations:

$$N\_{\rm iz} = N\_{\rm Tz} + N\_{\rm R} - N\_{\rm Kz} \tag{5}$$

where: *NTz*, *NR*, *NKz* – internal power of the turbine, reciprocating expander and combined-cycle compressor, respectively.

Specific power of the system was defined as follows:

$$N\_{\rm jz} = \frac{N\_{\rm iz}}{G\_{\rm K}} \tag{6}$$

Analysis of the results obtained for the Brayton-Diesel system model allows to draw following conclusions:

8 Will-be-set-by-IN-TECH

Fig. 11. Thermal efficiency of a combination of a gas turbine and reciprocating expander (configuration as in Fig. 10) as a function of specific power and TIT (T3). Numerical values

Fig. 12. Increase of a specific power output of a system with a reciprocating expander in comparison to the specific power output of a simple cycle gas turbine as a function of gas

An interesting solution could be also supplying fuel to the reciprocating expander, which would also require combustion process there. This case is currently a subject of further

The Brayton-Diesel system can also be used for combined heat and power applications. Mixture of exhaust gas and air at the outlet of the heat exchanger has a temperature in range

Internal power of the combined cycle *Niz* was calculated according to the rules and relationships given in Miller (1984), by binding it to the parameters of the thermal process

where: *NTz*, *NR*, *NKz* – internal power of the turbine, reciprocating expander and

*Njz* <sup>=</sup> *Niz GK*

Analysis of the results obtained for the Brayton-Diesel system model allows to draw following

*Niz* = *NTz* + *NR* − *NKz* (5)

(6)

shown in the chart denote gas turbine pressure ratios.

of 170–330◦C, so it can be used as a source of process heat.

turbine pressure ratio and TIT (T3).

investigation.

conclusions:

according to equations:

combined-cycle compressor, respectively.

Specific power of the system was defined as follows:

Fig. 13. Process diagram of a Brayton-Kalina system: C – compressor, T – turbine, G – generator, CC – combustion chamber, ST – steam turbine, K – Condenser, HRSG – heat recovery steam generator, HTHE – high temperature heat exchanger, LTHE – low temperature heat exchanger.


While selecting design parameters of the Brayton-Diesel system also the economical analysis is very important. Increasing TIT and using very large heat exchanger could prove too expensive when compared to the benefits resulting from the higher efficiency.

### **2.3 Brayton-Kalina**

The Kalina system is a variant of the Organic Rankine Cycle (ORC). It utilises a mixture of water and ammonia as a working agent. Ratio between ammonia and water may be changed, depending on the process, which enables adjusting boiling and condensation temperatures. The Kalina cycle is based on the Rankine cycle with added distilling and absorption components (Fig. 14). Its main advantage is already mentioned ability to adjust boiling and

Electricity obtained from renewable energy sources (water, wind, tidal)

Electricity generation Hydrogen Gas Turbine Cycle

form and simplicity of waste heat utilisation.

**3. Hydrogen gas turbine**

industrial areas.

world Gretz (1995); Kaya (1995).

H2 production – water electrolysis (solid polymer electron membrane SPEM)

Fig. 15. Configuration of a hydrogen-based power generation system

of combustion waste management which still remains unsolved.

**3.1 Review of the WE-NET system concept with hydrogen turbine**

Liquid H2 combustion – H2GT

Pure O2 production AIR

Gas Turbines in Unconventional Applications 131

gas turbine unit. The other possibility is recovering exhaust gas heat downstream from the gas turbine (as shown in the Fig. 14). Performance of this configuration may be optimised, with one of the variables being materials used for construction of the main heat exchanger. A Brayton-Stirling plant is able to achieve relatively high efficiency values. In a system built some years ago and utilising a Rolls-Royce RB211 gas turbine installation of a Stirling engine allowed to increase the power output by 9 MW (from the original 27.5 MW). This increased the total efficiency Poullikkas (2005) to 47.7%. And advantage of such a solution is its compact

The issue of environment pollution caused by human social and economical activities in recent years, has caused an increase of interest in environment-friendly (zero-emission) technologies of energy generation and utilisation. This approach is commonly seen as a way to solve global environmental problems that have recently become more and more intense. Environmental footprint of classic energy generation technologies based on fossil fuels – which are currently in use – is an argument for developing alternative technologies for energy generation. Power generation based on release of chemical energy of combusted hydrocarbon fuels is one of the main causes of the biosphere degradation. Combustion of fossil fuels (coal, lignite, oil, gas) causes an environmental footprint: emission of combustion to the atmosphere or the problem

An increase of CO2 concentration in the atmosphere intensifies the greenhouse effect and thus influences climate changes. Moreover, SO*<sup>x</sup>* and NO*<sup>x</sup>* emissions caused by many industry processes, especially combustion used in energy industry, causes acid rains falling in

Hydrogen as a clean fuel is a subject of interest for many research institutions all over the world. National and international research projects aimed at utilising hydrogen generated by renewable energy sources (e.g. wind or solar) have been proposed in various parts of the

The Japanese World Clean Energy Network programme using hydrogen conversion (WE-NET) started in 1993 as a part of larger "New Sunshine Program", whose main goal was to develop environment-friendly energy generation technologies that could meet constantly increasing energy demand and at the same time allow to solve arising environmental

H2

liquefaction plant Liquid H2

Liquid H2 Liquid H2

storage

Fig. 14. Process diagram of a Brayton-Stirling system. C – compressor, T – turbine, G – generator, CC – combustion chamber

condensation temperatures of the working agent, allowing to adjust entire cycle to the variable temperature of the heat source. Changing boiling and condensation temperatures during plant operation provide an additional degree of freedom when compared to a traditional steam cycle (including ORC solutions). Variable water content in the ammonia solution allows to optimise operational parameters of the cycle and thus improve the efficiency which may be close to the Carnot cycle.

Working agent properties allow to decrease the temperature difference in reference to the exhaust gas temperature in the heat recovery steam generator (the pinch point effect in evaporator is missing).

The Kalina cycle has a number of features making it competitive against the Rankine cycle (both "classic" and organic):


The licence for construction of Brayton-Kalina systems (process illustrated in Fig. 14) has been owned by General Electric since 1993. While further technology development was declared and a commercial power plant based on it was supposed to be completed by 1998, in fact the programme was suspended.

### **2.4 Brayton-Stirling**

The combination of Brayton and Stirling cycles can have different configurations. Heat to the Stirling engine may be delivered in two different points of the cycle. One possibility is heat transfer through a heat exchanger installed directly inside the combustion chamber of a 10 Will-be-set-by-IN-TECH

CC

2 3

Fig. 14. Process diagram of a Brayton-Stirling system. C – compressor, T – turbine, G –

condensation temperatures of the working agent, allowing to adjust entire cycle to the variable temperature of the heat source. Changing boiling and condensation temperatures during plant operation provide an additional degree of freedom when compared to a traditional steam cycle (including ORC solutions). Variable water content in the ammonia solution allows to optimise operational parameters of the cycle and thus improve the efficiency which may be

Working agent properties allow to decrease the temperature difference in reference to the exhaust gas temperature in the heat recovery steam generator (the pinch point effect in

The Kalina cycle has a number of features making it competitive against the Rankine cycle

2. It has lower (by some 40%) space requirements than a corresponding system with a steam

3. Exhaust steam pressure is higher than ambient pressure so it is not required to create

4. The Kalina cycle can be easily optimised for changed ambient conditions by adjusting

The licence for construction of Brayton-Kalina systems (process illustrated in Fig. 14) has been owned by General Electric since 1993. While further technology development was declared and a commercial power plant based on it was supposed to be completed by 1998, in fact the

The combination of Brayton and Stirling cycles can have different configurations. Heat to the Stirling engine may be delivered in two different points of the cycle. One possibility is heat transfer through a heat exchanger installed directly inside the combustion chamber of a

C

fuel

5

flue gas

air

generator, CC – combustion chamber

close to the Carnot cycle.

evaporator is missing).

turbine.

(both "classic" and organic):

working agent composition.

programme was suspended.

**2.4 Brayton-Stirling**

1

0

G

1. It may produce 10–30% energy more than a Rankine cycle.

vacuum in the condenser – this enables to shorten the start-up time.

T

flue gas

G

4

Fig. 15. Configuration of a hydrogen-based power generation system

gas turbine unit. The other possibility is recovering exhaust gas heat downstream from the gas turbine (as shown in the Fig. 14). Performance of this configuration may be optimised, with one of the variables being materials used for construction of the main heat exchanger. A Brayton-Stirling plant is able to achieve relatively high efficiency values. In a system built some years ago and utilising a Rolls-Royce RB211 gas turbine installation of a Stirling engine allowed to increase the power output by 9 MW (from the original 27.5 MW). This increased the total efficiency Poullikkas (2005) to 47.7%. And advantage of such a solution is its compact form and simplicity of waste heat utilisation.
