Technical and economical feasibility of the Rankine compression gas turbine (RCG)

Applied Thermal Engineering 26 (2006) 413 420 www.elsevier.com/locate/apthermeng Technical and economical feasibility of the Rankine compression gas turbine (RCG) H. Ouwerkerk *, H.C. de Lange Eindhoven University of Technology, P.O. Box 513, 5600 MB Eindhoven, The Netherlands Received 20 January 2005; accepted 9 June 2005 Available online 8 August 2005 Abstract The Rankine compression gas turbine (RCG) is a new type of combined cycle, i.e. combined steam and gas turbine installation, that returns all shaft power on one free power turbine. The novelty of the RCG is that the steam turbine drives the compressor of the gas turbine cycle. This way, the turbine of the gas turbine acts as a free power turbine. With its free power turbine the possible field of application of the RCG is mechanical drives. The RCG can be designed with components that can all be referred to as existing technology, which makes the RCG robust and technologically feasible. Thermodynamic calculations show that a thermal efficiency of about 40% is realistic. This is higher than simple cycle gas turbines, and equal to gas turbines with a recuperator. The calculations also show that the specific power of an RCG is up to twice as high than that of both simple cycle and recuperative cycle gas turbines. Finally, economical assessments show that the extra investments of an RCG compared to a simple cycle have an expected payback time of 2 4 years. This makes the RCG economically appealing, but further study is necessary to obtain more exact figures on the economical feasibility. Ó 2005 Elsevier Ltd. All rights reserved. Keywords: Combined cycle; Gas turbine; Mechanical drive; Free power turbine 1. Introduction For mechanical drives (driving pumps, natural gas compressors, etc.), internal combustion engines or simple cycle gas turbines with a free power turbine are employed. Internal combustion engines and gas turbines each have their own merits. To choose between the internal combustion engine and the gas turbine, many considerations have to be taken into account. This paper does not discuss these considerations because they are case dependant. For that reason this paper is restricted to the comparison of gas turbine based layouts. Due to ever increasing costs of fossil fuels and the awareness of the impact on the environment of burning fossil fuels, it is sought to decrease fossil fuel consumption. Various * Corresponding author. Tel.: +31 40 2475410; fax: +31 40 2475399. E-mail address: h.ouwerkerk@tue.nl (H. Ouwerkerk). technological developments are employed to lower the fuel consumption and emissions of mechanical drive gas turbines, such as high temperature materials and advanced combustion technologies. Also, efforts are being made to employ a recuperative cycle [8]. For micro-gas turbines (shaft power up to 100 kw) the recuperative cycle is already successfully employed [13,14]. This paper discusses a feasibility study of a new innovation with the goal to lower the fuel consumption of gas turbine based mechanical drives. In power stations, the combined cycle (combined gas turbine and steam turbine installation) is successfully employed to generate electricity at high efficiencies. Many analyses have been made of various combined cycles such as [4 6]. Recently an innovation was proposed [1,2], to employ the combined cycle also in mechanical drives. The innovation involves a new type of combined steam and gas turbine installation (combined cycle), 1359-4311/$ - see front matter Ó 2005 Elsevier Ltd. All rights reserved. doi:10.1016/j.applthermaleng.2005.06.007

414 H. Ouwerkerk, H.C. de Lange / Applied Thermal Engineering 26 (2006) 413 420 Nomenclature c g c g st g t g th c p h _m P Q fuel r ratio of specific heats:1.4 isentropic compressor efficiency isentropic steam turbine efficiency isentropic turbine efficiency thermal efficiency constant pressure specific heat [kj/kg] enthalpy [kj/kg] mass flow [kg/s] power [kj/s] heat added by burning of fuel [kj/s] compression ratio T temperature [K] Subscripts a air g gas ex exhaust gas l liquid st steam stack point where the exhaust gasses leave the steam generator which returns all shaft power by means of one free power turbine. This means that this combined cycle installation is going to be able to operate at rapidly changing speeds, and give torque when the load is standing still. For a combined cycle, this is unique. With this new technology, it will be possible to employ combined cycle installations in applications, where they could not have been employed before: mechanical drives, thus reducing fuel consumption by 15 25%. The invention is called the Rankine compression gas turbine (RCG). To assist the understanding of the RCG, first the two main existing types of combined cycle are shown in Fig. 1, the multi-shaft combined cycle (left) and the single-shaft combined cycle (right). Both make use of a compressor (C), combustion chamber, turbine (T), steam generator (waste-heat boiler), steam turbine (ST), condenser, water pump and generators. Both the combined cycles are shown driving generators, because that is their main application. Fig. 2 shows the layout of the new type of combined cycle, the Rankine compression gas turbine (RCG). The novelty of the RCG compared to existing combined cycles is, that the steam turbine (ST) drives the compressor (C) of the gas turbine cycle (Brayton cycle). Hence the name of the Rankine compression gas turbine: the compressor of the gas turbine cycle is powered by the steam turbine cycle (Rankine cycle). Otherwise, the RCG comprises the same components as the existing combined cycles. In Fig. 2 the turbine (T) is driving a load (L). Because the steam turbine (ST) drives the compressor (C), the turbine (T) acts as a free power turbine. In other words: the compression of the air required for the combustion chamber is powered by the waste-heat in the exhaust gasses of the turbine (T). Because the turbine (T) acts as a free power turbine, it will not only be able to drive an electrical generator, but also other loads, such as a pump or compressor (mechanical drive applications). To be able start the installation, an auxiliary burner (A) is fitted on the steam generator. The aim of this work is to investigate the technological and economical feasibility of the RCG. In order to assess the technological feasibility, design choices are made for the RCG to be applicable in mechanical drives. With a computer model, the efficiencies that can be reached with the chosen design choices are then determined. Then the characteristics of the RCG are determined. Finally, the economical feasibility is assessed. This is done by comparing the fuel reductions of the RCG to the investment costs. 2. Choice of components Because the RCG is meant for industrial applications, it will be in a much smaller shaft power range (2 MW up to 12 MW) than existing combined cycles. Also, the RCG will have to be robust, compact, and C Combustion chamber T G G ST Condenser Feedwater pump C Combustion chamber T ST Condenser G Feedwater pump Air intake Steam generator Exhaust Air intake Steam generator Exhaust Fig. 1. The multi-shaft combined cycle (l) and the single shaft combined cycle (r).

H. Ouwerkerk, H.C. de Lange / Applied Thermal Engineering 26 (2006) 413 420 415 Air-intake Condenser Gas turbine cycle C Combustion chamber ST Steam turbine cycle L T Feed water pump Steam generator Exhaust A Air-intake Fig. 2. The Rankine compression gas turbine (RCG). economic. Furthermore, dynamical behavior is a very important feature of industrial installations [3]. All this was recognized and accounted for in the preliminary design. In this section the choice of components is briefly discussed. 2.1. Compressor: centrifugal compressor and axial compressor As will be explained in Section 3, the gas turbine cycle of the RCG will typically have a pressure ratio of about 4. Because of this low pressure ratio, a centrifugal compressor can be employed for shaft powers up to about 3 MW. A centrifugal compressor is more robust and economic than an axial compressor. For higher shaft powers an axial compressor may be employed. 2.2. Steam turbine: impulse steam turbine In factory plants with central boilers, radial steam turbines are used to drive a variety of equipment in the shaft power range up to 15 MW. These radial steam turbines are in fact impulse steam turbines, because they have a reaction degree near to zero to be able to handle expansion ratios of up to 70, just like axial steam turbines. Only, radial steam turbines are much more compact, economic and robust than axial steam turbines. Recently, a new generation of radial impulse steam turbines has become available [10] with a turbine efficiency of up to 80%. These impulse steam turbines are very suitable for employment in an RCG-installation and can be considered as proven technology. 2.3. Steam generator (boiler) An RCG will need to have a steam generator that is compact and economic. So even though it will be at the cost of thermal efficiency, it is favourable to operate the boiler at relatively low pressure. This results in high temperature differences between exhaust gas and steam, so that a compact single stage boiler can be employed. 2.4. Condenser Robust industrial steam condensers are widely available in all sizes. Depending on the location and purpose of the RCG-installation, one will have to decide whether an air-cooled or water-cooled condenser is favourable. If enough cooling water is available, water-cooled condensers have the advantage of being more compact and economic. In the economical discussion of Section 5, the RCG is assumed to have an air-cooled condenser. 2.5. Feed water pump The water that is condensed by the condenser is pumped (and pressurized) into the steam generator by a feed water pump. Like the condenser, industrial feed water pumps are widely available. 2.6. Auxiliary burner To start an RCG, first the steam cycle has to be started. This way, the steam turbine will be powered up and will start to drive the compressor of the gas

416 H. Ouwerkerk, H.C. de Lange / Applied Thermal Engineering 26 (2006) 413 420 turbine cycle. Then the combustion chamber will be supplied with air and can be fired up. The steam cycle can be started with an auxiliary burner. The burner can be of the type that is used in industrial small-scale boilers (natural gas or oil fired). Note that during start up of the RCG, the power turbine can remain standing still until it gets enough hot gases from the combustion chamber to start driving the load. This means there is very little power needed to start an RCG. During start up of the RCG the only power consumption is that of the burner: the electrical fan for the burner air intake, pressurized natural gas (100 mbar) and ignition of the burner flame. 2.7. Combustion chamber and power turbine Both the combustion chamber and power turbine can be considered proven technology: gas turbine manufacturers have developed a large range of combustion chambers and turbines. For an RCG it would of course be favourable to employ an existing combustion chamber and power turbine of a gas turbine manufacturer. With the choices as described above, the RCG will be able to meet the preset requirements of being robust, compact and economic. 3. Thermodynamics For an RCG-installation in steady-state the power of the steam turbine has to be equal to the power consumed by the compressor of the Brayton cycle [1]: P compressor ½kWŠ ¼P steam turbine ½kWŠ. ð1þ The power consumed by the compressor is determined [8] with the following relation: T a P compressor ¼ _m a c c p;a 1. ð2þ g c And the power delivered by the steam turbine can be calculated [7] as follows: P steam turbine ¼ g st _m st ðh st;in h condenser Þ. ð3þ The amount of steam that can be generated by the steam generator follows from energy balance between the temperature drop of the exhaust gasses and the enthalpy rise of the water to steam in the steam generator [9]: r c 1 _m ex c p;ex ðt ex T stack Þ¼ _m st ðh st;in h l Þ. ð4þ Note that for a given turbine inlet temperature T TIT, the exhaust gas temperature after the power turbine T ex depends on the isentropic efficiency of the power turbine and on the pressure ratio r: a higher pressure ratio gives a lower exhaust gas temperature. Therefore, at steadystate, an RCG-installation will operate at a pressure ratio r for which both (1) and (4) are valid. The amount of shaft power generated by the RCGinstallation, is equal to the power of the power turbine [8]! P powerturbine ¼ _m g c pg g t T TIT 1 1 c 1 c : ð5þ r Finally, the thermal efficiency of an RCG-installation is determined by the ratio of the amount of fuel that is injected into the combustion chamber, and the power that is delivered to the load by the power turbine [1]: g th ¼ P powerturbine. ð6þ Q fuel With a computer model, comparative calculations were made of the RCG and the recuperative cycle [8]. For the RCG the above-mentioned choices of components were assumed, with the following properties (Table 1). For the recuperative cycle the pressure ratio (r) was optimized for maximum thermal efficiency. Furthermore, the following assumptions were made (Table 2). For the simple cycle [8], at turbine inlet temperatures higher than 1200 [K], it is not possible to optimize the pressure ratio (r) for maximum thermal efficiency, because the efficiency is ever increasing with increasing pressure ratio. It was chosen to compare the simple cycle at pressure ratioõs with a good balance between thermal efficiency and specific power [8] with the following assumptions (Table 3). Note that for the simple cycle higher isentropic compressor efficiency is assumed. This is an advantage for the simple cycle, but as the comparison will show, is also Table 1 Properties assumed for the RCG Ambient temperature 288 [K] g compressor 0.80 g turbine 0.85 g steam turbine 0.80 Boiler pressure 30 [bar] Steam temperature 773 [K] Condenser pressure/temperature 10 [kpa]/318 [K] Table 2 Properties assumed for the recuperative cycle Ambient temperature 288 [K] g compressor 0.80 g turbine 0.85 Recuperator efficiency 80% Table 3 Properties assumed for the simple cycle Ambient temperature 288 [K] g compressor 0.87 g turbine 0.85

H. Ouwerkerk, H.C. de Lange / Applied Thermal Engineering 26 (2006) 413 420 417 realistic: because the simple cycle has higher compression ratios an axial compressor is likely to be employed, while the RCG and recuperative cycle are likely to have a centrifugal compressor with their lower compression ratio. For both the steam generator and the recuperator, it is assumed that they do not result in backpressure for the power turbine. This gives a small advantage for the recuperative cycle, since a recuperator normally results in more backpressure than a steam generator. Fig. 3 shows the results regarding the obtainable thermal efficiency of the simple cycle, recuperative cycle and the RCG at varying turbine inlet temperature (TIT). The efficiency shown at a certain TIT is the efficiency of a certain installation with a turbine with that TIT as its maximum. The efficiencies of the simple cycle are much lower than of the recuperative cycle and the RCG, which of course was to be expected. The results show that both the RCG and the recuperative cycle can obtain efficiencies of about 30% up to about 45% in a range of realistic TITÕs. At the current maximum turbine entry temperature for an uncooled turbine, 1300 [K], they both rate a thermal efficiency of about 40%. At TITÕs higher than 1300 [K] the recuperative cycle has somewhat higher efficiencies than the RCG, but at TITÕs below 1300 [K] it is the RCG that has the highest efficiencies. It must be noted that the differences between the RCG and the recuperative cycle are small, and that the assumptions that were made are a little bit in favour of the recuperative cycle. These efficiencies were calculated, assuming modest component efficiencies, and without intercooling. So it can be concluded that the RCG will be an appealing alternative next to the recuperative cycle, when a higher efficiency than that of the simple cycle is preferred. thermal efficiency 0.5 0.45 0.4 0.35 0.3 0.25 0.2 0.15 0.1 0.05 0 1000 1100 1200 1300 1400 1500 1600 Turbine Inlet Temperature [K] RCG Rec. Cycle Simple Cycle Fig. 3. Thermal efficiency of the RCG, recuperative cycle and simple cycle at variable turbine inlet temperature. pressure ratio 25 20 15 10 5 0 1000 1100 1200 1300 1400 1500 1600 Turbine Inlet Temperature [K] RCG Rec. Cycle Simple Cycle Fig. 4. Pressure ratio of the RCG, recuperative cycle and simple cycle at variable turbine inlet temperature. Fig. 4 shows the corresponding pressure ratios of the compressor in the gas turbine cycle for the RCG, recuperative cycle and simple cycle. It can be seen that the pressure ratios of the simple cycle are much higher than those of the RCG and the recuperative cycle. This follows from the earlier discussed assumption to compare the simple cycle at pressure ratios with a good balance between thermal efficiency and specific power, because it is not possible to optimize for maximum thermal efficiency; for the simple cycle at maximum thermal efficiency, the specific power is equal to zero. The pressure ratio of the recuperative cycle is optimized for maximum thermal efficiency, and the pressure ratio of the RCG follows from the balance between the power of the steam turbine and the power consumption of the compressor. Most striking is, that the equilibrium pressure ratioõs of the RCG and the pressure ratioõs at optimum efficiency for the recuperative cycle are of the same magnitude. Further more, these pressure ratios can be realized with a centrifugal compressor. So the efficiencies of the RCG shown in Fig. 3, are those of a very robust RCG-installation: centrifugal compressor in the gas turbine cycle, impulse steam turbine and low-pressure boiler in the steam cycle. Of course existing combined cycles can achieve efficiencies of up to 54%. One could conclude that therefore it is no use introducing the RCG. That would be a false conclusion: the RCG is not meant to be a competitor of the existing combined cycles. The purpose of the RCG is, to make it possible to employ a combined cycle installation, where until now, this was not possible: the mechanical drive gas turbine applications. Further more, compared to the simple cycle and recuperative cycle, the RCG raises the specific power (kj/kg air/s) at the same turbine entry temperature (Fig. 5).

418 H. Ouwerkerk, H.C. de Lange / Applied Thermal Engineering 26 (2006) 413 420 Specific power [kw s/kg] 600 500 400 300 200 100 RCG Rec. Cycle Simple Cycle At equilibrium, the power from the steam turbine (ST) has to equal the power needed to drive the compressor (C). This results in a relatively low-pressure ratio for the compressor and thus the gas turbine cycle. Calculations show that for a robust RCG-system (no intercooling, etc.) the typical pressure ratio of the gas turbine cycle is about 3.0 5.5 with a thermal efficiency (shaft-power/fuel) of about 35% up to 45%, respectively. The relatively low-pressure index gives a high exhaust gas temperature. Because of this, there is a great temperature difference between the exhaust gas and steam, resulting in a compact and economic steam generator. 0 1000 1100 1200 1300 1400 1500 1600 Turbine Inlet Temperature [K] With these properties the RCG will be suitable for mechanical drive applications and possibly also ship propulsion applications. Fig. 5. Specific power of the RCG, recuperative cycle and simple cycle at variable turbine inlet temperature. Fig. 5 shows that the recuperative cycle and simple cycle have almost the same specific power, while the RCG has up to twice as much specific power. The basis for the large increase in specific power of the RCG compared to the recuperative and simple cycle is that the waste heat in the exhaust gasses is converted into shaft power. In the recuperative cycle the waste heat is also put to use, but it is employed to reduce the amount of fuel that is burned to reach the same turbine inlet temperature. In the RCG the waste heat is converted into shaft power that is employed to drive the compressor. So the turbine no longer has to drive to compressor and delivers all its power to the load. Note that the rise in specific power of the RCG compared to the recuperative and simple cycle is typical for all combined cycles, not just the RCG. 4. Technological feasibility To assess the technological feasibility we consider the main characteristics of the RCG: All shaft-power is available from one free power turbine (T). The RCG gives torque even when the shaft load (L) and power turbine (T) are standing still. To start the RCG, the steam cycle is to be powered by auxiliary firing (ordinary burner in the steam generator), the power turbine (T) and load (L) can remain standing still. As soon as the steam turbine (ST) and compressor (C) run at sufficient speed the combustion chamber can be started. Subsequently, when the power turbine (T) gives enough power, the load will speed up from zero. 5. Economical discussion To assess whether the RCG is economically appealing, a comparison is made with simple cycle gas turbines. Of course a lot of issues play a role in the economical feasibility, and they differ per application, user and country. In this study it was not possible to take all these matters into account. Although there are great differences in the prices of natural gas, and although for oil and gas companies natural gas is available at prices below the average, we still calculated with the average price: at this stage it is impossible to make exact calculations. Also it is likely that the reduction of CO 2 will give financial benefits in the future, but it is difficult to account for them; therefore they are neglected in the calculations. The goal is to roughly asses whether it is economically attractive to further develop the RCG. Because the assumptions are not very exact, of course the outcome of the calculations will also not be exact. However, all the systems are compared with the same assumptions. So all together the calculations will give a good idea whether the RCG is economically attractive, but for more exact figures, further study is necessary. Because the RCG is meant for the power range 2 12 MW, it is chosen to compare simple cycle and RCG-installations with a shaft power of 2.5 MW and 10 MW. The numbers that are shown in Table 4 for the two simple cycle installations are typical for actual installations. For the two RCG installations, the numbers are Table 4 Assumptions for comparison of the simple cycle and RCG Shaft power g th Investment costs ( ) 2.5 [MW] Simple cycle 0.30 1,400,000 RCG 0.35 1,800,000 10 [MW] Simple cycle 0.38 3,700,000 RCG 0.42 4,500,000

H. Ouwerkerk, H.C. de Lange / Applied Thermal Engineering 26 (2006) 413 420 419 the result of current market prices of the components that the RCG consists of, and of realistic estimates of the costs to assemble and realize the installation. Also, the condenser of the RCG is assumed to be aircooled. For 2.5 MW shaft power the extra investment costs of an RCG compared to the simple cycle are relatively higher than at 10 MW shaft power. This is mainly caused by the (impulse) steam turbine of the RCG. Although an impulse steam turbine is the most costeffective solution at these relatively low shaft powers, at 2.5 MW it is still more costly per MW shaft power than at 10 MW shaft power. When comparing the thermal efficiencies of the RCG and simple cycle, it shows that the gain in efficiency of an RCG at 2.5 MW is relatively higher than at 10 MW. The reason for this is that in general, gas turbines can be designed more efficient with increasing shaft power. This is because with increasing shaft power it becomes more cost efficient to employ a higher compression ratio (more stages) and better thermal resistant materials. Due to the lower thermal efficiency of the 2.5 MW simple cycle gas turbine, the exhaust gasses have a higher temperature, thus contain more energy to be utilized by a waste-heat steam cycle. Therefore the gain in thermal efficiency of an RCG compared to the simple cycle is the highest at the lower shaft power of 2.5 MW. Calculations were made assuming an average of 5.80/GJ for natural gas. This is the average price [11] for natural gas in Europe. Further more, an availability of 90% was assumed. This is not for maintenance reasons, it is just assumed that mechanical drives are not running full load all the time. With Table 4, the natural gas price and an availability of 90% the costs over the years can then be calculated (Fig. 6). Fig. 6 shows the total of the investment costs and fuel costs per kw, divided by the number of years. Comparing the two 2.5 MW installations to the two 10 MW installations, the curves of the 2.5 MW are above the curves for 10 MW. This is, of course, because the investment costs of the smaller installations are higher per kw and also their fuel costs per kw are higher because their thermal efficiency in general is lower than that of larger installations. Comparing the simple cycle to the RCG, Fig. 6 shows that both for 2.5 MW and 10 MW shaft power, the extra investments of an RCG-installation are paid back within about two years. Of course this two years pay back time is anything but exact. A lot of non-exact assumptions were made to make this calculation. But it can be stated that the expected payback time of the extra investments of an RCG installation compared to a simple cycle gas turbine is in the order of magnitude of a few years (two up to four years), which is economically very appealing. It is possible to design a RCG-installation that is smaller than 2.5 MW. The results are not shown in Fig. 6, but down to 500 kw they result in about the same curve as cumulative costs per kw per year x 1000 EUR 1.2 1.1 1 0.9 0.8 0.7 0.6 0.5 0.4 for 2.5 MW shaft power. Further more, it is not possible to design a RCG installation with more then 12 MW shaft power. This is because for more then 12 MW shaft power, a high-efficiency impulse steam turbine of over 6 MW is needed to drive the compressor. Such impulse steam turbines are not yet available. However, due to the free power turbine principle of the RCG, it is possible to implement multiple parallel impulse steam turbine and compressor units, supplying compressed air to the combustion chamber(s) of one larger power turbine. This type of installation was not studied yet. Further more, such an installation should then also be compared to an installation with one axial steam turbine. 6. Conclusions 1 2 3 4 5 years simple cycle 2.5MW RCG 2.5MW simple cycle 10MW RCG 10MW Fig. 6. Comparison of yearly costs of simple cycle and RCGinstallations. The Rankine compression gas turbine (RCG) is a new type of combined cycle. By arranging the existing combined cycle components in a new manner, the RCG has a (for combined cycles) unique load characteristic; all shaft power of the RCG is delivered by a free power turbine. With its free power turbine, the aimed field of application of the RCG is mechanical drives. The RCG can be designed in such a way that all necessary rotating components are commercially available. This makes the RCG technologically feasible because the components of the RCG can be referred to as existing and reliable technology. Thermodynamic calculations show that the RCG will offer thermal efficiencies of over 40% with modest components. In the mechanical drive shaft power range, this thermal efficiency is higher than

420 H. Ouwerkerk, H.C. de Lange / Applied Thermal Engineering 26 (2006) 413 420 simple cycle gas turbines and is equal to that of recuperative gas turbines. The calculations also show that the specific power of an RCG is very high; for TITÕs higher than 1300 [K], the specific power is about 60% higher than that of both simple cycle and recuperative cycle gas turbines. For TITÕs up to 1300 [K], the specific power can be up to a 100% higher. Because of the high specific power and due to the design choice to employ a cost-efficient impulse steam turbine, the RCG will offer relatively low investment costs. Economical assessments show that comparing the simple cycle gas turbine to the RCG, the extra investments of an RCG have an expected payback time of 2 4 years. These results show that the RCG is economically appealing, although further study is necessary to obtain more exact figures on the economical feasibility. However, considering the results of this study it is appealing to pursue the development of the RCG for mechanical drive applications in the shaft-power range of 2 12 MW. Experiments were conducted on a small-scale (100 kw) test set-up at the Technische Universiteit Eindhoven to prove the technological feasibility of the RCG [12]. Presently a small-scale (100 kw shaft power) prototype of the RCG is being realised in a laboratory of the Technische Universiteit Eindhoven. The prototype will be fully operational in the summer of 2005. After that, it is sought to realise a real-scale RCG. References [1] H. Ouwerkerk, Feasibility study of the Rankine compression gas turbine, Masters thesis (Dutch), Technische Universiteit Eindhoven, 2002. [2] H. Ouwerkerk, Patent filed PCT/NL03/00271, Steam and gas turbine installation, International filing date 9 April 2003, Priority date 10 April 2002. [3] H.A. van Essen, H.C. de Lange, Modelling and model based control of turbomachinery, Ph.D Thesis, Technische Universiteit Eindhoven, 1998. [4] T.J. Leo, I.P. Pérez-Grande, P. Pérez-del-Notario, Gas turbine turbocharged by a steam turbine: a gas turbine solution increasing combined cycle power plant efficiency and power, Appl. Therm. Eng. 23 (2003) 1913 1929. [5] T. Heppenstall, Advanced gas turbine cycles for power generation: a critical review, Appl. Therm. Eng. 18 (1998) 837 846. [6] F.M. Penning, H.C. de Lange, Steam injection: analysis of a typical application, Appl. Therm. Eng. 16 (1996) 115 125. [7] F. Dietzel, Dampfturbinen, Hanser, München, 1980. [8] H. Cohen, G.F.C. Rogers, H.I.H. Saravanamuttoo, Gas Turbine Theory, Longman Group Limited, England, 1996. [9] Y.A. Çengel, M.A. Boles, Thermodynamics: An Engineering Approach, McGraw-Hill, Singapore, 1989. [10] www.agkkk.de, website of impulse steam turbine manufacturer, Kühnle, Kopp & Kausch, Frankenthal, Germany. [11] www.gasinfo.be. [12] Abstract accepted and paper submitted for the International Symposium on Air breathing Engines (ISABE) 2005. [13] www.microturbine.com, website of micro turbine manufacturer, Capstone, US. [14] www.turbec.com, website of micro-turbine manufacturer, Turbec, Sweden.