Fluid Selectionfluid whr

1 Copyright © 2011 by ASME

Proceedings of ASME 2011 5th International Conference on Energy Sustainability & 9th Fuel Cell Science, Engineering and Technology Conference

ESFuelCell2011 August 7-10, 2011, Washington, DC, USA

ESFuelCell2011-54579

FEASIBILITY STUDY OF AN ORGANIC RANKINE CYCLE SYSTEM COUPLED TO A DIESEL ENGINE

Joseph DiCarlo Department of Mechanical & Industrial

Engineering University of Toronto

Toronto, Ontario, Canada

James S. Wallace Department of Mechanical & Industrial

Engineering University of Toronto

Toronto, Ontario, Canada

ABSTRACT A feasibility study has been conducted for an organic Rankine cycle (ORC) system coupled to the exhaust of a Diesel engine generator. The objective of this study was to determine the possible electrical generation of the ORC using the exhaust gas of a 78 kW diesel engine as its energy input. A thermodynamic model was developed to predict the possible electricity generation of the ORC. Using this model it was determined that the preferred working fluid for the ORC was R245ca. The calculated maximum ORC thermal efficiency was 14.3%. The net electrical power generated from the ORC was 5.36 kWe. The ORC would require no additional fuel and would not generate any additional emissions. The most cost-effective and simple means to develop a small packaged ORC system is to high volume production HVAC system components. An ORC system consisting of air conditioning system components, including a scroll expander, yielded a projected ORC efficiency of 10.7%, and a net electrical power of 4.02 kWe . The total capital cost of the ORC system was $2,140 CAD. Three engine usage scenarios were developed, with the time to payback of the initial capital cost of the ORC ranging from 12.8 to 0.58 years, depending on low or high usage.

INTRODUCTION Extensive research has been conducted to increase diesel engine efficiency, including combustion system development, turbocharging, friction reduction and many other refinements. Recovering exhaust waste heat energy is another way to enhance efficiency. Although

some exhaust energy is required to drive the turbocharger and facilitate reaction in exhaust aftertreatment sytems such as oxidation catalysts, diesel particulate filters, and selective catalytic reduction (SCR) systems, there is still energy that can be recovered. Organic Rankine Cycle (ORC) technology can be utilized to recover the available waste heat and provide electrical power to supplement the electrical output of a diesel engine powered generator set. The added electric power improves the overall efficiency of the generator set since no additional fuel would be required.

The objective of this study was to analyze the efficiency and economics of an ORC system coupled to the exhaust of a diesel generator. Engine data was obtained from a 78 kW diesel engine coupled to an engine dynamometer. Suitable working fluids were selected on the basis of their thermodynamic properties, environmental impact, and safety. Thermodynamic analysis was then carried out for each working fluid to determine the efficiency and possible electrical power production for the ORC system. An optimum ORC working fluid was chosen for this application. Major system components were selected to make the ORC physically realizable. An economic analysis was conducted encompassing the major component costs. The initial cost, and the time to system payback, based on the cost of electricity (that would otherwise need to be purchased) saved by utilizing the ORC was then compared to the projected gain in efficiency to evaluate the economic feasibility.

Proceedings of the ASME 2011 5th International Conference on Energy Sustainability ES2011

August 7-10, 2011, Washington, DC, USA

ES2011-54579



NOMENCLATURE

Symbol Physical Characteristic Units

m

air Mass flow rate of air into the Diesel engine

kg/s

m

fuel Mass flow rate of fuel into the Diesel engine

kg/s

m

e Mass flow rate of exhaust from Diesel engine

kg/s

m

r Mass flow rate of refrigerant in the ORC

kg/s

Tc Temperature of the cooling water °C Te Temperature of the Diesel engine

exhaust °C

T1,T2,T3,T4

ORC state temperatures °C

Tv Temperature at the saturated vapor state

°C

Ta Temperature of the Diesel engine exhaust gas

°C

Tb Temperature of exhaust gas after heat exchanger

°C

Tcr Critical Temperature °C TTDevap Terminal temperature difference

of the evaporator °C

TTDcond Terminal temperature difference of the condenser

°C

Pboiler Boiler operating pressure bar Psat Saturation pressure bar P1,P2,P3,P4 ORC state pressures bar Pcr Critical Pressure bar h1,h2,h3,h4 ORC state specific enthalpies kJ/kg ha Specific enthalpy of the Diesel

engine exhaust gas kJ/kg

hb Specific enthalpy of exhaust gas after heat exchanger

kJ/kg

hv Specific enthalpy at the saturated vapor state

kJ/kg

sv Entropy at the saturated vapor state

kJ/kg-K

s1,s2,s3,s4 ORC state specific entropies kJ/kg-K

s4s Isentropic specific entropy at state 4

kJ/kg-K

v1 Specific volume at state 1 of the ORC

m3/kg

x Vapor quality M Molar mass g/mol wpump Specific work required by the

pump kJ/kg

wturb Specific work output from the turbine

kJ/kg

wgen Specific work generated by the electric generator

kJ/kg

wmotor Specific work required by motor to power pump

kJ/kg

wnet Net specific work produced by the ORC

kJ/kg

qin Specific thermal energy input to working fluid

kJ/kg

qout Specific thermal energy output from working fluid

kJ/kg

ηpump Pump efficiency % ηturb Turbine efficiency % ηORC Overall efficiency of the ORC

system %

ηgen ORC generator efficiency % ηmotor Motor efficiency %

Q

in Rate of heat transfer from the engine exhaust gas to the working fluid

kW

W

net Rate of net work produced by the ORC

kW

ORC BACKGROUND ORC technology is not new. It is estimated that about 30 commercial ORC plants were built before 1984 with an output of 100 kW [4]. ORC plants are typically used in four main capacities: waste heat recovery, biomass power plants, geothermal power plants, and solar thermal power plants. Leslie et al. [5] describe a 5.5 MW ORC system developed as a bottoming cycle for gas turbine powered gas pipeline compressors. GE and ECOS Ltd. developed a 300 kW ORC bottoming cycle used with a 7.2 MW biogas facilty [6]. These are both waste heat recovery applications [bottoming cycles] similar to the diesel generator set application investigated in the present study. The first geothermal power plant in Germany, at Neustadt-Glewe, is an example of an ORC geothermal system. It uses hot water at approximately 98oC to produce 210 kW of electricity [7]. Cooper and Wallace [3] carried out a feasibility study of a 200 kW solar thermal ORC power generation system. Although the thermal efficiency of ORC systems is low due to the small temperature difference available, the cost of electricity production can be reasonable if the capital cost of the equipment is low. A recent cost breakthrough was the recognition that an air conditioning chiller could be operated with minor modifications as a power generation system [21]. High volume production of these system components substantially reduces ORC system cost and enhances the economic case for power production from relatively low grade waste heat. The present study is predicated on availability of inexpensive components.

ORC WORKING FLUID SELECTION Selection of the working fluid is an essential design step for an ORC. Four criteria were utilized in the selection: thermodynamic properties, environmental impact, thermal stability, and safety.


Figure 1. T-s diagram of working fluids [12]

Thermodynamic stability - The shape of the saturated vapor curve in temperature-entropy (T-s) diagram of a working fluid is one of the most important factors in determining suitability for ORC use. Fluids with a negative saturated vapor curve slope (wet fluids) become a two-phase mixture of liquid and vapor during expansion. This is undesirable because liquid droplets can damage turbine blades. Superheating is the usual remedy for steam plants, but the added cost and complication of superheating are not viable for ORC systems. Working fluids with a vertical slope (isentropic

fluid) or a positive slope (dry fluid) are suitable for ORC use. The critical temperature of the working fluid is also very important. The critical temperature of the working fluid should approximately match the temperature of the system energy source. All other things being equal, this will maximize the thermal efficiency of the cycle. Figure 1 shows that the critical temperature of water is demonstrably higher that the critical temperatures of refrigerants. In addition, the T-s curve for water is much wider than the T-s curve for a refrigerant. The greater latent heat of vaporization for water (1989.8 kJ/kg at 20 bar) compared to R245fa , a representative ORC working fluid (110.42 kJ/kg at 20 bar), implies that ORC working fluids will require a higher mass flow rate than water to utilize a given system heat input.

Environmental Impact - The environmental impact of a working fluid can be judged on its Ozone Depletion Potential (ODP) and its Global Warming Potential (GWP). ODP is a measure of a fluid’s ability to deplete stratospheric ozone relative to R11, which has an ODP of 1. GWP is an index describing a fluid’s relative ability to trap radiant energy compared to CO2. Table 1 summarizes environmental data of common organic working fluids [16].

Table 1: Environmental data for selected organic fluids [16]



Chlorofluorocarbons (CFCs) and hydrochlorofluorocarbons (HCFCs) can affect both stratospheric ozone and climate change, whereas hydrofluorocarbons (HFCs) can affect climate change, but not stratospheric ozone [15].

Thermal stability - The thermal stability of the working fluid in an ORC application is very important. Organic fluids usually suffer chemical deterioration at high temperatures. In their study of zero ODP working fluids, Angelino et al. [17] determined experimentally that most of the organic fluids suffer visible chemical deteriorations at temperatures in excess of 300°C. Thus, the maximum temperature of an ORC system is limited based on the selected working fluid. Furthermore, the freezing point of the fluid has to be lower than the lowest ambient temperature the ORC system will encounter.

Safety - The working fluid must be non-flammable, non-corrosive and have a low toxicity to minimize the human health and environmental impact of any leakages.

Preliminary selection of working fluids - A preliminary selection of 5 working fluids was made based on the optimal characteristics given above. The most critical factor was ODP. All refrigerants with an ODP > 0 were eliminated from the analysis. It was decided that the working fluid must be either isentropic or dry. Of the organic fluids that met the above criteria, those with the 5 highest critical temperatures were chosen for the analysis. The fluids with the highest critical temperatures, R134a, R245ca, R245fa, R236ea, and R236fa, will allow a greater temperature difference between the boiler and the condenser, and this will result in a greater efficiency. Thermodynamic properties of these refrigerants are given in Table 2.

Table 2: Properties of the five acceptable working fluids [18]

Fluid M (g/mol)

Tcr (°C)

Pcr (bar)

Saturation Dome Shape

R134a 102.3 101.1 40.6 isentropic R245ca 134.1 174.4 39.3 dry R245fa 134.1 154.1 32.0 dry R236ea 152.0 139.3 35.0 dry R236fa 152.0 124.9 32.0 dry

ORC SIMULATION MODEL A simulation of the ORC was set up using

Microsoft Excel software. Figure 2 shows a simple ORC system incorporating a boiler, expander, condenser, and pump. The model also takes into account a generator coupled to

Figure 2: ORC Model

the expander, and a motor that runs the pump. The exhaust from the engine enters the ORC at state a, transfers thermal energy to the organic working fluid in the boiler heat exchanger, and exits at state b. For this model, the condenser is modeled as a shell and tube heat exchanger with water as the coolant. However, an air-cooled condenser is also feasible. Regardless of which means of cooling is used as the basis for the model, the thermal energy rejected to the environment during the condensation process will be equivalent. Figure 3 shows a T-s diagram that describes the ORC model for a dry working fluid. Since the model does not incorporate superheating, state 3 of the working fluid is taken to be the saturated vapor condition. Mago et al. [19] concluded that the specific irreversibility increases and the second-law efficiency decreases with superheating in an ORC system with a dry working fluid.

Figure 3: Generic T-s diagram for the ORC Model



GOVERNING EQUATIONS

The equations that constitute the model and the required input parameters are presented below.

Model input parameters: The model requires

values as inputs for the following parameters: m

air,

m

fuel , Tc, Te, Pboiler, TTDcond, TTDevap, T1, ηpump, ηturb, ηgen, ηmotor.

Model equations: State 1

P1 = Psat(T1) h1,v1,s1 = hf,vf,sf (T1)

State 2 P2 = Pboiler

wpump = v1(P2 P1)

pump

T2 ≈ T1 h2 = h1 + wpump

State 3 T3 = Tv (Pboiler) P3 = Pboiler

h3 = hv (T3) s3 = sv (T3)

State 4 P4 = P1 s4s = s3

x4s = s4 s sf

sfg

if x4s ≤ 1:

h4s = hfg*x4s + hf else: h4s = Linear Interpolation (s4s @ Pcond

superheater vapor)

wturb = ηturb(h3 −h4s) h4 = h3 − wturb

Engine exhaust parameters

m

e m

air + m

fuel State a

Ta = Te ha = h(Ta)

State b Tb = T3 + TTDevap

hb = h(Tb) State c hc = hf (Tc) State d Td = T1 - TTDcond hd = hf (Td)

ORC system parameters (Including energy transfers as work and heat) Refrigerant mass flow rate

m

r m

e(hb ha )

(h2 h3)

qin = h3 – h2

qout = h4 – h1 wgen = gen wturb

wmotor = wpump

motor

wnet = wgen - wmotor

ORC wnet

qin

Q

in qin m

r

W

net wnet m

r The most important model outputs are the

refrigerant mass flow rate, m

r, the thermal energy input

rate, Q

in, and the rate at which net work is produced,

W

net. The refrigerant mass flow rate is an important factor in determining the required component sizes. The thermal energy input rate is the energy the ORC recovers from the exhaust gas. The rate at which net work is produced is the ORC’s electric power production.

MODEL INPUTS Table 3 provides an overview of the model inputs, and their values for the 5 selected organic working fluids (Table 2) used for this analysis. These values were selected as they are expected to approximate actual operating conditions. The condenser temperature (T1) is 30°C. At this temperature, the use of an air-cooled condenser is feasible [21]. The boiler pressure was set at 20 bar. This is taken to be an acceptably low pressure that will allow relatively high vapor temperatures. The engine exhaust gas data was obtained a 78kW diesel engine operating at 75% of its maximum power output. Thus, for the purposes of this analysis, the engine is producing:

78 kW * 0.75 = 58.5 kW

MODEL OPERATION The National Institute of Standards and Technology (NIST) standard reference database [11] is used to obtain fluid thermodynamic properties for the Excel ORC model. The exhaust gas from the engine was modeled as air, and the enthalpies of air at various temperatures were obtained using an online air table [20].



Table 3: Excel ORC Model Inputs

Temperatures and Pressures Tc (°C) 12.0 Condenser TTD [TTDcond] (°C) 10.0 Exhaust gas Temperature [Te] (°C) 425.0 Boiler TTD [TTDevap] (Tb-T3) 20.0 Boiler Pressure [Pboiler] (bar) 20.0 T1 (°C) 30.0 Massflows Engine air [mair] (kg/s) 0.130 Engine fuel [mfuel] (kg/s) 0.004 Engine Load (%) 75.0 Efficiencies Pump Efficiency [ηpump] (%) 65.0 Turbine Efficiency [ηturb] (%) 85.0 Generator Efficiency [ηgen] (%) 90.0 Motor Efficiency [ηmotor] (%) 90.0

The Excel ORC model can be used to obtain the desired outputs using the following iterative procedure: 1) Select desired working fluid. 2) Enter the model inputs, as given in section 5.3.

3) Use NIST to obtain isothermal data at T1. Input to Excel worksheet.

4) Use NIST to obtain isobaric data at Pboiler. Input to Excel worksheet.

5) The model Solves for Pcond. Use NIST to obtain isobaric data at Pcond. Input to Excel Worksheet.

6) Use NIST to obtain Saturation Curve data. Input to Excel worksheet.

Appendix A contains further detail on using the NIST database to obtain thermodynamic properties. After completing the above 6 steps, the Excel model generates the outputs and the T-s diagram.

RESULTS The results obtained from the Excel ORC analyses for each of the 5 working fluids can be found in Table 4. A description of the critical results follows, including T-s diagrams, condenser pressure, electricity production and thermal energy transfer T-s Diagrams - The Excel model uses the outputs to develop T-s diagrams of the working fluids. These diagrams are extremely useful as they aid in understanding the fluid properties and their impact on heat transfer and efficiency. The T-s diagrams for each working fluid can be found in figures 4 through 8.

Table 4: Results of Excel ORC model analyses for the 5 organic working fluids

OUTPUTS R245ca R245fa R134a R236ea R236fa

Temperatures and Pressures

Refrigerant max T3 (°C) 136.7 122.0 67.5 111.7 101.5

Expander exit T4 (°C) 61.3 54.4 30.0 54.6 47.0

Condenser P1 (bar) 1.22 1.79 7.70 2.44 3.21

Electrical Power

Electricity input to motor [Wmotor] (kW) 0.32 0.38 0.45 0.43 0.50

Net Electricity generated [Wnet] (kW) 5.36 5.11 3.15 4.73 4.36

Thermal Energy Transfer

Heat transfer from exhaust gas to ORC [Qin] (kW) 37.5 39.5 46.9 40.9 42.3

Condenser Heat transfer [Qout] (kW) 31.5 33.7 43.3 35.6 37.3

Massflows

Refrigerant [mr] (kg/s) 0.14 0.16 0.25 0.20 0.24

Efficiencies

ηORC (%) 14.3 12.9 6.7 11.6 10.3


Figure 4: T-s diagram R245ca

Figure 5: T-s diagram R245fa

Figure 6: T-s diagram R134a

There are several critical features to note in these diagrams. Firstly, the critical temperatures of the 245 series refrigerants, figure 4 for R245ca and figure 5 for R245fa, are demonstrably higher than the critical temperatures of the other working fluids. The critical temperature of R134a is especially low, at 101.05°C, as shown in Figure 6. The T-s characteristics of R236ea and R236fa can be seen in Figure 7 and 8, respectively. The

.

Figure 7: T-s diagram R236ea

Figure 8: T-s diagram R236fa

critical temperatures are worse than the 245 series, but superior to R134a, for this specific ORC application. The refrigerant properties determine the boiler temperature. At the 20 bar boiler design pressure, R245ca has a saturation temperature of 136.73°C, whereas R134a has a saturation temperature of 67.5°C.

ORC condenser pressure – Figure 9 shows that the ORC system condenser pressure varies, depending on

Figure 9: ORC condenser pressure for each

working fluid


Figure 10: Wnet vs. Winput (pump parasitic power) for various working fluids

which working fluid is chosen. A 30°C condenser temperature was used as an input for all five working fluids to allow for the use of an air-cooled condenser. The condenser pressure is determined by the saturation pressure of the working fluid at this temperature. The saturation pressure of R134a is 7.7 bar, which may be too high for standard air-cooled condensers However, the saturation pressures of the other four refrigerants are more reasonable. In general a saturation pressure above atmospheric pressure is desirable to prevent air leakage into the condenser.

ORC net electricity production - The function of the ORC system is to produce electricity from the diesel engine’s exhaust energy. However, Figure 2 shows that some electricity is consumed within the ORC system to drive the pump. The best working fluid should maximize the net electricity produced. The results for net electrical work and required electrical pump work are displayed in Figure 10. R245ca has the greatest net electrical work produced (5.36 kWe), and the smallest required electrical input (0.322 kWe). R245fa has similar results, although it is slightly worse. The working fluid with the poorest performance was R134a. It would produce a net electrical output of 3.15 kWe and would require an electrical input of 0.449 kWe. This is due to its low critical temperature, and subsequent low operating temperature. Thus, the choice of working fluid has a dramatic impact on the ORC system’s net electrical power output.

Thermal energy transfer - The thermal energy input to the ORC is the waste heat recovery from the diesel engine exhaust, which occurs in the boiler heat exchanger. The ORC heat rejection occurs in the condenser, as the thermal energy remaining in the working fluid at the outlet of the expander is transferred to the environment. The resultant thermal energy transfers from the model results can be found in Figure 11. This chart shows the thermal power input to the working fluids during the boiling process and heat rejection from the working fluids during the condensation

Figure 11: Predicted thermal energy transfers in

the ORC

process. R134a has the greatest thermal energy input from the exhaust gas at 46.9 kWth. This is due to the lower boiler temperature of the R134a refrigerant in the heat exchange process. The temperature difference between the exhaust gas and the working fluid is the driving factor for heat transfer between the two fluids. R245ca has the lowest thermal energy input from the exhaust, due to its greater boiler temperature. 37.5 kWth are transferred from the exhaust gas to the R245ca working fluid. What is most important is the difference between Qin and Qout. This difference is the potential energy that can be used to produce electrical work, which is the purpose of the ORC. The difference between these two values is greatest for R245ca, at 6 kWth. It is important to note the heat rejection from the condenser. The large magnitude of the heat rejection, relative to the thermal energy input, demonstrates the limitations of trying to generate electricity from a low temperature heat source. Although the net power taken from the exhaust may be small in magnitude, if the ORC can be developed with acceptable economics, the project may be a reasonable investment.

Boiler heat exchanger temperature profile- As mentioned above, the boiler heat exchanger (Evaporator) functions based on a Terminal Temperature Difference (TTD). An evaporator terminal temperature difference of 20 °C was used for this analysis, i.e. the temperature of the exhaust gas at the heat exchanger outlet is 20 °C higher than the temperature of the working fluid at the heat exchanger outlet. Figure 12 shows the R245ca heat exchanger temperature profile, where the heat exchanger outlet is the 100% heat transfer point for both fluids. Note that the working fluid goes through two distinct phases as it is heated up in the exhaust gas heat exchanger. Approximately 60% of the heat transfer is made as the working fluid is in the process of going from subcooled liquid to saturated liquid. The remainder of the heat exchange occurs as the working fluid goes from a saturated liquid to a saturated vapor. The magnitudes of the thermal energy transfers can be seen in Table 5.


Figure 12: Evaporator heat exchange profile for

R245ca

This is markedly different from a simple Rankine cycle utilizing water, where only a small portion of the heat exchange is made when the working fluid is in the subcooled region. The reason for this can be seen in Figure 1. This chart shows the T-s diagram of R245fa and water superimposed upon each other. This is an excellent figure for visualizing the differences between the two fluids. Water has a much larger difference between saturated liquid and saturated vapor specific enthalpy than the refrigerant does, so for water the heat of vaporization portion of the heat transfer dominates. The temperature of the fluid does not increase during this two-phase region due to the phenomena of latent heat of vaporization. The additional thermal energy added to the working fluid is utilized to facilitate boiling, rather than increasing the temperature.

Mass flow rate and efficiency - The Excel ORC model also predicts mass flow rates and efficiencies. The refrigerant mass flow rate and density are important factors to consider when determining the size of the components required. Figure 13 shows the mass flow rates required and Figure 14 the cycle thermal efficiency for each of the five refrigerants. R245ca has the lowest mass flow rate (0.14 kg/s) and has the greatest efficiency (14.3%), when compared to the other refrigerants. It would thus require the smallest components (i.e. pump, expander, exhaust gas heat exchanger), and would be the most efficient. Table 5: Thermal energy transfers in the

evaporator

Process Thermal Energy Transfer (kJ/kg)

State 2 -> Sat. liquid 20 bar

159.11

Sat. liquid 20 bar -> State 3

113.35

Total (State 2 -> State 3) 272.46

Figure 13: ORC working fluid mass flow rates

Figure 14: ORC efficiency for each fluid

Working fluid recommendation - For the given inputs and based on the criteria presented earlier, the working fluid that would be best for this ORC application is R245ca. It produces the greatest net electrical power, has the highest efficiency, has the smallest mass flow rate requirement, and operates at an acceptable condenser pressure.

PREDICTED INCREASE IN GENERATOR PERFORMANCE As noted earlier, the engine inputs to the ORC are taken from a diesel engine operating at a power output of 58.5 kWe. The ORC system with R245ca working fluid would produce a predicted net electrical output of 5.36 kWe. The ORC system output would increase the total electrical output to 63.86 kWe, an increase of 9.16%. The gross electrical efficiency increase can also be calculated, under the assumption that the Diesel engine is approximately 40% efficient. In this case the thermal energy input to the Diesel engine is 146.25 kW.

The

gross electrical efficiency of the combined system is: (58.5kW 5.36kW )

146.25kW 0.437

This represents an absolute increase in gross electrical efficiency of 3.7%, from 40% to 43.7%, or a 9.25%



Table 6. Diesel generator efficiency with and without the ORC system

Diesel engine

Diesel engine w/ORC

Absolute improvement

Percentage improvement

Electrical Power Generated

58.5 kW 63.9 kW 5.36 kW 9.16%

Gross Electrical Efficiency 40% 43.7% 3.7% 9.25%

Table 6: Impact of scroll expander on ORC electrical power generation

ORC w/ original expander ORC w/ scroll expander % Change

Expander efficiency 85% 65% -23.53%

Wnet 5.36 kW 4.02 kW -25.00%

ηORC 14.3% 10.7% -25.17% increase in efficiency when compared to the baseline. For convenience these values are tabulated in Table 6.

COMPONENT SELECTION Major system components, including expander, condenser, pump and evaporator, were selected to facilitate economic analysis. A central home air conditioner could provide components for a small ORC system. A typical outdoor condensing unit consists of a compressor, a condenser, and an expansion valve. The compressor and motor can be converted into an expander and generator respectively. The condenser can be used directly as an ORC system condenser. Thus starting with an air-conditioning condensing unit, only a pump and an evaporator would be needed to complete the ORC system.

Expander selection – In developing a packaged ORC system from HVAC equipment, Brasz et al. converted an air conditioning centrifugal compressor into a turbine [21 with only minor modifications. However, only very large air conditioning systems utilize a centrifugal compressor. Modern central air conditioning systems utilize high efficiency scroll compressors [25]. Quoilin et al. converted a scroll compressor into an expander, and through experimental testing demonstrated an isentropic efficiency of 65% [26]. Thus, it should be possible to effectively utilize a condensing unit having a scroll compressor. A Goodman GSX13 condensing unit with a maximum power input of 10 kW was selected. It is available for $895 CAD.

Evaporator selection - Since the ORC is designed with the exhaust of a diesel engine as its heat source, an exhaust gas heat exchanger must be used. A cross-flow heat exchanger, Polar model 30, was selected. It has a price of $895 CAD. It is not specifically designed as an evaporator but is of a suitable size and material construction as to provide a reasonable cost estimate. Pump selection – The pump must have sufficient capacity to pressurize the working fluid to 20 bar, the evaporator pressure. The pump selected was the Stuart Turner Monsoon 20 Bar single impeller pump. It has a price of $350.00 CAD [30].

Thermodynamic analysis revisited - Usage of the scroll expander requires a slight modification of the thermodynamic analysis. The isentropic turbine efficiency used for the initial analysis was 85%, as shown in Table 3. The isentropic efficiency of the scroll expander is taken to be 65%, based on the work of Quoilin et. al [26]. Table 6 shows revised results factoring in this decrease in expander efficiency. This 23.53% decrease in expander efficiency decreases the predicted electrical power production of the ORC by 25%, from 5.36 kWe to 4.02 kWe. Figure 15 shows the impact of the decrease in expander efficiency on a T-s diagram. As expected, the only difference occurs between states 3 and 4 (the expansion process). For the 65% case, state 4 is at a higher enthalpy level, and thus less energy is utilized as work, and more energy is released in the condensing process. Ideally, one would want to minimize thermal energy loss during the condensation stage. Although this 25% decrease in electrical power is substantial, it is important to note that the resulting 4.02 kWe of electrical power will be produced with no additional fuel and emissions from the diesel engine.

Figure 15: The impact of expander efficiency



Table 7: Cost of ORC components

Unit Cost ($ CAD) GSX13 Condensing unit 895.00 Polar Heat Exchanger Model 30 895.00 20 bar pump 350.00 Total Capital Cost $2,140.00

Table 8: Electricity prices in Ontario [31]

Regulated Electricity Prices Cost ($CAD/kWh)

Energy charge - first 1000 kWh/month

0.058

Energy charge - after first 1000 kWh/month

0.067

Debt Retirement 0.007 Distribution 0.0186 Transmission 0.0102 System operation and regulation 0.0062 ECONOMIC ANALYSIS Table 7 presents a summary of the ORC system component costs. The total predicted capital cost of the ORC system is $2,140 CAD. Note that the price of the GSX13 Air Conditioner includes refrigerant charging. For the purposes of this analysis, it was assumed that the cost difference between the air conditioning refrigerant and the ORC working fluid would be negligible when compared to the system component costs. The total capital cost estimate was obtained from the retail prices of an air conditioner, an exhaust gas heat exchanger, and a pump. It is thus considered to be a conservative cost estimate. If the ORC were packaged together and mass produced, the price would likely drop substantially, and retail prices for individual components could be avoided. Given the simplicity of an ORC system of this size, maintenance costs were assumed to be negligible. In addition, there is no fuel cost for an ORC, since it operates without fuel requirements. The breakdown for the price of electricity in Toronto can be found in Table 8. This pricing data can be used to estimate the value of money saved by utilizing the electricity generated by the ORC, rather than purchasing electricity from a utility.

Daily usage generator - If the Diesel engine generator were used every day, one could consider the hours per day usage when attempting to determine the time to payback for the ORC system. This relationship has been plotted in Figure 16. There is an approximate threshold daily usage, ranging from 4 to 8 hours per day, above which the time to payback slowly approaches the <1 year range, and below which the time to payback rapidly approaches the >10 year range. Within this 4 to 8 hours/day threshold range, the time to payback varies from 3.66 to 1.83 years.

Figure 16: Daily usage vs. time to payback

Generator usage scenarios - An economic analysis was conducted under the following three scenarios: low usage, medium usage, and high usage of the generator. Further details on the economics of each of the three scenarios can be found in Appendix D. Scenario 1: Low generator usage - The generator is used mostly for emergency backup power. It is used 8 hours a day, 1 day a week, for 52 weeks a year. A summary of the power generation can be found in Table 9. The money saved, and the time to payback on the original capital cost, can be found in Table 10.

Table 9: Low generator usage electricity

production

Power generatio

n (kW)

Operating

ime (hrs/yr)

Yearly energy

generated (kWh/yr)

Monthly energy

generated (kWh/month)

4.02 416 1,672.32 139.36

Table 10: Low generator usage economic

summary

Total monthly cost saved($/month) 13.94

Total yearly cost saved ($/yr) 167.23

TIME TO PAYBACK (YEARS) 12.8 In this scenario, the time to payback is 12.8 years.

Scenario 2: Medium generator usage - The generator is used 8 hours a day, 5 days a week, 44 weeks a year. This scenario represents supplementary power production for a multi-residential or business applications. It is also representative of generator usage on construction sites. The power generation can be found in Table 11. These inputs yield the economic data shown in Table 12.



Table 11: Medium generator usage electricity production

Power

generation

(kW)

Operating

time

(hrs/yr)

Yearly

energy

generated

(kWh/yr)

Monthly

energy

generated

(kWh/month)

4.02 1,760 7,075.2 589.6

Table 12: Medium generator usage economic summary

Total monthly cost

saved($/month) 58.96

Total yearly cost

saved ($/yr) 707.52

TIME TO PAYBACK (YEARS) 3.02 Note that the time to payback is 3.02 years.

Scenario 3: Heavy generator usage - The generator is used to aid in power generation of a small village. This scenario is also representative of CHP generator usage for a hospital, apartment building, etc. It is used for 24 hours a day, 7 days a week, 52 weeks a year. The power generation can be found in

Table 13. The economic summary for this scenario can be found in

Table 14.

Table 13: Heavy generator usage electricity production

Power generation

(kW)

Operating

time

(hrs/yr)

Yearly

Energy

generated

(kWh/yr)

Monthly

energy

generated

(kWh/month)

4.02 8,736 35,118.72 2,926.56

Table 14: Heavy generator usage economic summary

Total monthly cost saved ($/month) 310.00

Total yearly cost saved ($/yr) 3,719.94

TIME TO PAYBACK (YEARS) 0.58 The time to payback in this scenario is 0.58 years.

Summary Based on the above economic analyses for the

three scenarios, it is evident that the ORC system is only a feasible solution for the medium-to-heavy engine usage

scenarios. Although eventually all scenarios will cover the initial capital cost, the capital cost of the ORC system may not be justifiable when factoring in the 12.8 year payback time for the low generator usage scenario.

CONCLUSIONS This study explored the feasibility of coupling an

ORC system to the exhaust of a Diesel engine generator to generate additional electricity. A model was developed to predict the additional electricity generation possible by utilizing an ORC coupled to the exhaust of a 78 kW Diesel engine generator. A preferred organic fluid was selected using this model in conjunction with a preliminary study on working fluids. This preliminary study also factored in human health and environmental impacts of the working fluid. Components were selected to make the system physically realizable. The impacts that the real components would have on the system performance were factored in. An economic analysis was conducted to determine the time to payback for utilizing this system.

The following main conclusions were developed:

1. The working fluid for this system which met all objectives, and yielded the highest efficiency and net power output, was R245ca.

2. The calculated maximum ORC thermal efficiency was 14.3%. The net electrical power generated was 5.36 kWe.

3. Utilization of high-volume production air conditioning system components is the most cost-effective and simple means to develop small packaged ORC systems.

4. Converting an air conditioning system with a scroll compressor to a ORC with a scroll expander yielded a projected efficiency of 10.7%, and a net electrical power of 4.02 kWe. This is due to the 65% isentropic efficiency of a scroll expander.

5. This electrical power would be produced with no additional fuel requirements or emissions from the diesel engine.

6. The estimated total capital cost of the ORC system was $2,140 CAD.

7. Three engine usage scenarios were developed, with the time to payback of the initial capital cost ranging from 12.8 to 0.58 years.

Utilization of an ORC system to generate electricity from exhaust waste heat is a feasible approach to improving the electrical efficiency of a Diesel engine generator. However this feasibility is dependent upon the generator usage. For low generator usage, the time to payback may not be acceptable based on an economic analysis. For high generator usage, the time to payback is relatively short and the initial capital cost would be acceptable to improve the overall efficiency of the generating system.



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