Design study of configurations on system COP for a combined ORC (organic Rankine cycle) and VCC (vapor compression cycle)

Energy 36 (2011) 4809e4820

Contents lists available at ScienceDirect

Energy journal homepage: www.elsevier.com/locate/energy

Design study of configurations on system COP for a combined ORC (organic Rankine cycle) and VCC (vapor compression cycle) Hailei Wang*, Richard Peterson, Tom Herron School of Mechanical, Industrial, & Manufacturing Engineering, 204 Rogers Hall, Oregon State University, Corvallis, OR 97331, USA

a r t i c l e i n f o

a b s t r a c t

Article history: Received 12 November 2010 Received in revised form 4 April 2011 Accepted 15 May 2011

The study introduced a novel thermally activated cooling concept e a combined cycle couples an ORC (organic Rankine cycle) and a VCC (vapor compression cycle). A brief comparison with other thermally activated cooling technologies was conducted. The cycle can use renewable energy sources such as solar, geothermal and waste heat, to generate cooling and power if needed. A systematic design study was conducted to investigate effects of various cycle configurations on overall cycle COP. With both subcooling and cooling recuperation in the vapor compression cycle, the overall cycle COP reaches 0.66 at extreme military conditions with outdoor temperature of 48.9
C. A parametric trade-off study was conducted afterwards in terms of performance and weight, in order to find the most critical design parameters for the cycle configuration with both subcooling and cooling recuperation. Five most important design parameters were selected, including expander isentropic efficiency, condensing and evaporating temperatures, pump/boiling pressure and recuperator effectiveness. At the end, two additional cycle concepts with either potentially higher COP or practical advantages were proposed. It includes adding a secondary heat recuperator in the ORC side and using different working fluids in the power and cooling cycles, or so-called dual-fluid system. Ó 2011 Elsevier Ltd. All rights reserved.

Keywords: Organic Rankine cycle (ORC) Thermally activated cooling Vapor compression cycle COP Waste heat R245fa

1. Introduction With increasingly world-wide environmental concerns about global warming due to carbon emission associated with fossil fuel consumptions, renewable energy sources such as solar, wind, geothermal have gained significant attention internationally as clean energy sources. Moreover, the soaring cruel oil price witnessed in recent years reminds the world again that it cannot solely depends upon fossil fuels to meet the ever-increasing energy needs. Alternative energy sources have to be part of the global energy solutions, which again indicates the importance of the renewable energy sources such as solar, wind and geothermal. Because of their renewable nature, they are also considered as free energy. Likewise, waste heat is considered as clean and renewable energy in some countries, since there is no additional direct carbon emission and they are free energy otherwise would be wasted. ORC (Organic Rankine Cycles) as a mature technology for low temperature power generation has been around for at least three decades. Barber and Prigmore conducted some pioneer work related to ORC using solar thermal energy in mid 70s [1,2]. It

* Corresponding author. Tel.: þ1 541 713 1354; fax: þ1 541 758 9320. E-mail address: [email protected] (H. Wang). 0360-5442/$ e see front matter Ó 2011 Elsevier Ltd. All rights reserved. doi:10.1016/j.energy.2011.05.015

re-gained significant interests in recent years due to environmental and economical concerns. It remains to be a strong candidate for converting low-to-medium grade heat into power due to its technology maturity and relatively high thermal conversion efficiency. It has been successfully demonstrated and studied using renewable energy sources such as solar [3e7], geothermal [8e12] and waste heat [13e26] in recent years. Besides power generation, ORC can be used for many other applications. An application that has attracted particular interests in southern Europe recently is to use sunlight to produce fresh water. The studies were focusing on using solar-power ORCs to drive desalination systems based on RO (reverse osmosis) process [27e39]. This is a rather important field as fresh water remains as another critical source on earth with ever-growing population. Besides, fresh water is generated using sunlight, which has no direct impact on the environment. The power generated in the ORC is directly used to drive a high pressure pump in the RO system. The outcome is fresh clean water. Another potentially important application for ORC could be in environmental control or air-conditioning field. Most residential air conditioning systems typically use VCC (vapor compression cycle) to create cool air for comfort. It requires electricity in order for a motor to drive the compressor and generate cooling. Cooling is achieved by consuming high-grade electricity, given the fact that

4810

H. Wang et al. / Energy 36 (2011) 4809e4820

Nomenclatures A COPc COPs h _c m _p m Ppump PR Q_ boi Q_ eva Tboil Tcond Tevap U _ com W _ exp W

hcom hexp hp

Erecp

Heat transfer area (m2) Cooling cycle coefficient of performance Overall system coefficient of performance Fluid enthalpy (kJ/kg) Cooling cycle mass flow rate (kg/s) Power cycle mass flow rate (kg/s) Pump outlet pressure (kPa) Pressure ratio between expander inlet and outlet pressures Boiler heat input (kW) Evaporator cooling capacity (kW) Boiler outlet temperature (
C) Condensing temperature (
C) Evaporating temperature (
C) Overall heat transfer coefficient (W/m2-C) Compressor work input (kW) Expander work output (kW) Compressor isentropic efficiency Expander isentropic efficiency Power cycle conversion efficiency Primary power recuperator heat transfer effectiveness

electricity is generally produced by some sort of thermal energy. A typical home use air conditioner or heat pump on the market now has the SEER (seasonal energy efficient ratio) of 13, which corresponds to the COP (coefficient of performance) around 3.43. If taking into account the thermal-to-electricity conversion efficiency of 25%e30% for a typical power generation system, the actual heatto cooling COP of a typical A/C or heat pump unit is between 0.85 and 1. As it is shown in Fig. 1, the demand for electricity surges during summer hours as people seek for more comfort environment, especially in hot and humid regions. It not only causes problems such as power outage, but also dramatically increases energy cost for homeowners during the summer in order to run a typical air conditioner. In contrast, the demand for natural gas during summer hours drops substantially. This is the fundamental rationale of developing and improving thermally activated cooling technologies, not to mention using free and renewable thermal sources such as solar, geothermal and various waste heat streams. Although there are thermally activated cooling technologies such as absorption cycles, they are generally used for large scale

industrial applications as absorption chillers. The COP is generally low for single stage absorption cycle (COP < 0.7) using LiBreH2O pair [40e45]. To go to multiple stages such as triple-effect systems, their cost increases dramatically and sometime is prohibitive due to stringent material requirement to prevent corrosion at high temperature [46e51]. There are small-scale absorption systems under developments using ammonia-water, generally referring to generator-absorber heat exchange absorption cycle or GAX cycle [52e59]. However, the system is relatively complicated and not yet proven to be high COP and economically viable. Besides, there are some additional risks associated with ammonia, including toxicity, corrosion and extreme high pressure. Because of that, it has not yet been widely recognized. Desiccant cooling has attracted more interests in recent years, given its thermally activated cooling nature and being able to precisely control and maintain low humidity of air [60e74]. In order to maintain low humidity for some applications, current VCC systems are required to reduce air temperature below its dew point temperature to condense out the moisture. Air is then heated to a comfort temperature before delivered for space cooling. This significantly reduces the efficiency of the compressor due to higher temperature lift, and thus requires significantly more power from the motor. This alone reduces both the COP and cooling capacity of air conditioners. Therefore, it requires significant more power for an A/C or heat pump to operate in hot and humid climate. In contrast, desiccant cooling systems can achieve this with less energy input besides using low-grade thermal energy. Significant progress has been made to improve the performance of desiccants [75e77]. Recent studies have shown the promise of using composite desiccant materials, which has higher specific moisture adsorption capacity at lower regeneration temperature [78e80]. The thermally activated cooling concept of combining an ORC with a VCC has not yet been given enough attention such as the absorption cycles and the ORC-RO systems for fresh water production. There was one similar study combining an organic Rankine cycle with a vapor compression refrigeration cycle published recently [81]. It uses one piston design for the expander and compressor versus scroll expander and compressor design for the current study. The concept proposed in the current study combines an ORC power cycle with a VCC cooling cycle to form a thermally activated cooling system. The shafts of the expander in the ORC and compressor in the VCC are directly coupled to reduce two-way energy conversion losses. Although the proposed ORC-VCC has some potential advantages over other thermally activated cooling systems in terms of performance and simplicity, a comprehensive comparison study is needed to be effective and conclusive. Nevertheless, compared to absorption and desiccant cooling cycles, the ORC-VCC system has the flexibility of outputting power when cooling is not needed. This applies to applications that continuous waste or renewable heat is available throughout all seasons. In hot summer, all heat can be converted to cooling, while only part of the heat is converted to cooling in spring and fall. No heat is converted to cooling in winter. When cooling is not needed, all the thermal energy can be converted to power and sent to the grid. 2. Cycle configurations

Fig. 1. Year Round Natural gas and Electricity Demand.

Most previous studies related to ORC were concentrating on either cycle performance analysis or fluid selections. It is the focus of this study to optimize cycle configurations. EES (Engineering Equation Solver) was used to model the combined ORC and VCC system. The fundamental equations of state based on Helmholtz energy were used to calculate the thermodynamic properties such as enthalpy and entropy at each cycle state. The analysis started

H. Wang et al. / Energy 36 (2011) 4809e4820

cases that the pressure ratio kept the same, the fluid superheat remained the same. For cases with higher pressure ratios, the level of superheat changed accordingly. Following the basic cycle, the model was used to investigate the effects of subcooling of the condensate in the cooling cycle. After that, cooling cycle recuperation was taken into consideration. Throughout this modeling work, pressure drop across each component was neglected. A single working fluid in both the cooling and power cycles was assumed.

13

Boiler Expander

Qin,hot 14

Power Recuperator

12

11 15

Compressor

7

4811

8

1) Basic cycle Condenser 10

6

Qout

Evaporator Pump

Qin,cool Receiver

mp mc

4

1

Expansion Valve

Fig. 2. Schematic flow diagram of the basic combined ORC-VCC cycle.

with a basic cycle, which has evaporating and condensing temperatures of 18
C and 67
C, respectively. The reason for such high condensing temperature and temperature lift (difference between condensing and evaporating temperatures) is that the unit was designed based upon stringent military conditions e outdoor air temperature of 48.9
C (120
F). As for commercial A/C unit, the design outdoor temperature is typically 35
C. Another reason for such high condensing temperature was due to size and weight constrains on the system design. higher condensing temperature makes the overall system more compact and light-weighted for potential portable applications. It should be noted that, if the system were designed based on the commercial conditions, the COP even for the basic cycle would be significantly higher. The fluid vapor temperature coming out of the boiler was kept at 190
C. For

Fig. 2 shows the schematic flow diagram of the basic ORC-VCC cycle. Fluid from the condenser (1) passes through the expansion valve (4) and is vaporized in the evaporator (6). It is then recompressed in the compressor (7), and condensed to liquid in the condenser. It is known that the compressor is typically driven by a motor in a typical vapor compression cycle, this ORC-VCC cycle uses an integral power generating portion of the cycle to produce the shaft work required by the compressor. The power generating portion pumps liquid from the condenser to high pressure and delivers it to a boiler (13) where it is vaporized and superheated. The vapor then passes through an expander (14) where work is produced to drive the compressor. The spent vapor passes through a recuperator (15) which uses sensible heat in the expander exhaust to preheat liquid entering the boiler. The addition of the recuperator typically boosts the overall cycle COP by 30%e40%. This improvement is one of the features that make this ORC-VCC cycle attractive. It should be noted that the state point numbers are not consecutive. They are chosen to be consistent with subsequent diagrams containing more components. The diagram shows a receiver at the outlet of the condenser. This small vessel not only ensures that liquid exits the condenser at saturation and maximizes the heat transfer from the condenser, but also it partially functions as a fluid reservoir. It helps reduce system sensitivity to fluid charge amount. Fig. 3 shows the combined cycle on a pressureeenthalpy diagram for the assumed working fluid, R245fa. Although it is not

Fig. 3. Pressure-enthalpy diagram of the basic combined ORC-VCC cycle.

4812

H. Wang et al. / Energy 36 (2011) 4809e4820

Table 1 Model Inputs for the Basic Cycle. Parameter P_pump Value

T_boil

_ W

T_cond T_evap PR hexp hcom erecp Q_ eva

3200 kPa 190
C 67
C

18
C

8

75% 80%

85% 5.3 kW

an ideal working fluid for compact vapor compression cycles due to its low vapor pressure, R245fa is commonly chosen as the working fluid for organic Rankine cycles given its desirable thermodynamic and thermal transport properties such as low boiling temperature and relatively high thermal conductivity. It is ideal for low to medium temperature applications, e.g. waste heat recovery, solar thermal and geothermal power generation. In addition, it is nonflammable, non-ozone-depleting, non-corrosive and has very low toxicity. The pressureeenthalpy diagram includes the saturation dome, cycle process lines and constant property lines for temperature, entropy, and quality. The lower box is the vapor compression cooling portion of the cycle, and the upper box on the diagram represents the power generation portion of the cycle. The main operating conditions of the cycle are listed in Table 1. The P-h diagram is insightful because the specific enthalpy change in a process is represented by the horizontal distance between endpoints of the process. For instance, the specific heat absorbed in the evaporator is proportional to the length of line 4e6, and the specific work done by the compressor is horizontal distance covered by the line sloping line 6e7. The cooling cycle COP of the cooling cycle is then the ratio of these two horizontal distances. Thus the graph provides a visual indication of the cycle performance. The effect of the recuperator is shown by the dashed lines. Heat released from the exhaust stream in process 14e15 is absorbed by the feed stream in process 10e12, assuming perfect thermal insulation. The heat input by the boiler is indicated by process 12e13. From the graph, it can be seen that the recuperator provides more than a quarter of the overall power cycle heat input. A set of equations used to evaluate the cycle performance are listed in the following.

hp ¼ _ exp Q boi

(1)

_ exp ¼ m _ p ðh13  h14s Þhexp W

(2)

_ p ðh13  h12 Þ Q_ boi ¼ m

(3)

The cooling cycle COP is defined as:

Q_ eva _ com W

(4)

_ c ðh7S  h6 Þ _ com ¼ m W

(5)

_ c ðh6  h4 Þ Q_ eva ¼ m

(6)

_ com ¼ W _ exp W

(7)

COPc ¼

hcom

The overall gross COP of combined cycle is defined as:

COPg ¼

Q_ eva Q_

(8)

boi

In order to calculate the net COP, power consumption by the pump and fan motors was taken into account. Both pump and fans were planned to use brushless DC motors, which generally have high motor efficiency. A rotary piston pump was identified for its high efficiency and small footprint. After combining the pump and its motor efficiencies, the typical power consumption for the pump assembly is 1.5 times of the flow work. Regarding fan power, a correlation was developed between fan power and the UA of individual heat exchanger such as the condenser based on the motor efficiencies provided by relevant fan manufacturers. U is the overall heat transfer coefficient and A is the heat transfer area of

Fig. 4. Schematic flow diagram of the combined ORC-VCC cycle with subcooling.

H. Wang et al. / Energy 36 (2011) 4809e4820

4813

Fig. 5. Pressure-enthalpy diagram of the combined ORC-VCC cycle with 85% subcooling.

the heat exchanger. Knowing the pump and fans power, the net COP of the combine cycle is defined as:

Q_ eva

COPn ¼ Q_ boi þ

_ pump þ W _ W fans

(9)

0:25

where 0.25 was to assume 25% of generator efficiency. In the basic cycle, the model shows that the ORC side thermal conversion efficiency is 14.7% and the COP for the vapor compression cooling cycle is 3.67, which results in the combined cycle gross COP of 0.54.

2) Basic cycle with subcooling At the end of the condensation process, the fluid is still significantly above ambient temperature and can be cooled further with ambient air. Although from the practical standpoint the fluid needs to be slightly subcooled in order to prevent pump from cavitating in the power cycle, this is undesirable from a performance standpoint since it increases the amount of heat that must be added back to the fluid in the boiler. In the cooling cycle, however, subcooling is beneficial. Fig. 4 shows a schematic flow diagram of the combined ORC-VCC cycle with a subcooler added between the receiver and

Fig. 6. Schematic flow diagram of the combined ORC-VCC cycle with cooling recuperation.

4814

H. Wang et al. / Energy 36 (2011) 4809e4820

Fig. 7. Pressure-enthalpy diagram of the combined ORC-VCC cycle with cooling recuperation.

Table 2 Performance Summary of Various Cycle Configurations. Cycle Configuration

Basic Cycle

With Subcooler

With Cooling Recuperator

Gross COP

0.54

0.63

0.66

the expansion valve. This provides subcooling to only the fluid flowing through the cooling portion of the cycle. Fig. 5 shows the peh diagram of the cycle with subcooling. Without subcooling, nearly 35% of the fluid flashes off in the expansion valve without contributing to cooling. By cooling the fluid 85% of the temperature difference between saturation and ambient conditions, the quality after the expansion valve is reduced to 0.23. This increases the specific cooling provided by the fluid in the evaporator and leads to the cycle gross COP of 0.63.

Subcooling will also likely reduce system weight along with increasing efficiency, since weight is one of the targets in this design study. In the cooling portion, the increase in efficiency reduces the size of the compressor and decreases the size of condenser needed. A smaller compressor also means that a smaller power cycle is needed e expander, boiler, recuperator, and condenser all get smaller in proportion to the increase in efficiency. The only cost is an additional heat exchanger about 1/8 the size of the condenser. Clearly subcooling is very desirable for the combined ORC-VCC cycle. 3) Basic cycle with subcooling and cooling recuperation To take the concept of subcooling a step further, a cooling recuperator (or suction line heat exchanger) is added to the cycle to further subcool the liquid entering the expansion valve using the

Fig. 8. Schematics diagram of the weight model.

H. Wang et al. / Energy 36 (2011) 4809e4820

4815

Fig. 9. Major component weight for each cycle configuration.

3. Parametric study Apparently the overall cycle COP increases, as the cycle becomes more sophisticated and more components involved. Table 2 is a summary of the cycle COP for different cycle configurations. In practical, however, the most efficiency system will not always be designed and built. In order to freeze to a particular design, other design constrains need to be taken into account such as size, weight, and cost targets for the overall system. In the following study, the basic cycle with subcooling and cooling recuperation was further investigated with system weight as one of the design targets. A weight model as illustrated in Fig. 8 was developed for the system. For rotating components such as the expander and compressor, the model was scaled to their mass flow rates. The weights of heat exchangers, on the other hand, were developed correlations with their UAs. The rest of system had fixed weight. Although it is desirable to have precise weight models, their

90

0.8

0.7

85

0.7

0.5 Gro

0.4

Ne

ss

t

C

75

P CO

70

90 G ro

ss C

85

OP

0.6 Ne t C

COPS

80 OP

System Weight, kg

0.8

0.6 CO PS

compression. The recuperator provides ample superheat to avoid any dilution problems and potentially can remove the accumulator.

80 OP

0.5

75

Weig ht

0.4

System Weight, kg

cold vapor exiting the evaporator. Its schematic flow diagram is shown in Fig. 6. The effect of the cooling recuperator is to further reduce the enthalpy of the fluid entering the expansion valve (highlighted line between points 2 & 3 in Fig. 7). This increases the cooling effect of the fluid in the evaporator from the subcooling case (cross hatched highlight) by the amount of heat transferred in the recuperator (solid highlight). The effectiveness of the recuperator is assumed to be 85%. Unlike the subcooler, the cooling recuperator does affect the work required by the compressor. By heating the vapor entering the compressor, the density of gas is reduced so that less mass is compressed for the same work. This increase in work reduces the efficiency benefit produced by the additional subcooling. Still, the recuperation raises the gross COP to 0.66. While the efficiency benefit of the cooling recuperator is modest, there is a practical advantage that should be mentioned. A common requirement for mechanical compressors is that the vapor is significantly superheated to avoid dilution of the lubricant by absorbed working fluid, and/or an accumulator is added before the compressor to further protect the compressor. This is especially important for R245fa which is prone to condensation during

70

Weight

0.3 0.5

0.6

0.7

0.8

65 0.9

Isentropic Efficiency Fig. 10. Effect of expander isentropic efficiency on overall system COPs and weight.

0.3 63

64

65

66

67

68

69

70

71

65 72

Temperature, °C Fig. 11. Effect of condensing temperature on overall system COPs and weight.

0.8

90

0.7

85

0.7

85

COP

0.6

G ross

0.5

Ne t CO P

80

75

Weight

0.4

0.3 12

13

14

15

16

17

18

19

0.8

90

0.7

85

COPS

0.5

Ne

CO

P

80

OP

tC

75

0.4 Weight 0.3 2000

3000

4000

Net COP

75 t Weigh

70

65 0.6

0.7

0.8

0.9

Fig. 14. Effect of power recuperator effectiveness on overall system COPs and weight.

Figs. 10 to 14 show the effect of expander isentropic efficiency, condensing temperature, evaporating temperature, pump/boiling pressure and power recuperator effectiveness on the combined cycle gross COP, net COP and system weight. As the Fig. 10 shows, the cycle performance strongly depends on expander isentropic efficiency. As expander isentropic efficiency increases, the COPs increase proportionally. At the same time, weight is reduced with increasing efficiency. The dashed line indicated the value assumed in the Cycle Configuration section. The condensing and evaporating temperatures shown in Figs. 11 and 12, on the other hand, produce a trade-off between COPs and weight. Reducing condensing temperature, for instance, increases COPs but also increases weight. The condensing temperature chosen in the Cycle Configuration section gives a weight near the minimum value while maintaining good COPs. The graph for evaporating temperature shows COPs could be increased by raising the evaporating temperature without significantly increasing weight. Another advantage of higher evaporating temperature is that latent cooling is reduced. However, 18
C is already substantially higher than typical values used in vapor compression systems. Further increasing evaporating temperature could result in a very big evaporator. Two other significant factors in system performance are pump or boiling pressure and power recuperator effectiveness. Their effects are plotted in Figs. 13 and 14. According to Fig. 13, there is an optimal pump pressure at which COPs are highest while weight is lowest. The optimal pump pressure is actually close to 3200 kPa, which is the value used in the previous section. Both COPs and weight could be improved by increasing pumping pressure, if isentropic efficiency

System Weight, kg

relative values are most important for the following parametric study. Fig. 9 shows the typical component weight for each of the three cycle configurations. This study focused on a system with designed cooling capacity ðQ_ eva Þ of 5.3 kW, which determined the mass flow rates of the _ p Þ cycles. For some applications such as _ c Þ and power ðm cooling ðm waste heat recovery, it is desirable to recover as much as waste heat as possible. However, those conditions may not always match the conditions that highest efficiency of the ORC and/or VCC occurs. Higher boiling pressure or temperature for the ORC fluid, although generally raises the ORC-VCC cycle efficiency, it causes the waste heat stream leaves the system at higher temperature. Therefore less thermal energy in the waste heat stream is utilized. Besides for a given expander, its efficiency generally drops as pressure ratio increases (see Fig. 15). Higher boiling pressure could actually lead to lower cycle performance as shown in Fig. 13 without other penalties such as increased weight. Alternatively, higher cycle COP can be maintained with sacrifice of system weight by designing the expander with multi-stage expansion. In order to characterize the combined system further at different operating conditions with system weight, pump and fan power taken into account, a series of parametric studies were conducted.

ss

0.5

80

Effectiveness

Fig. 12. Effect of evaporating temperature on overall system COPs and weight.

o Gr

Gross C

0.3 0.5

Temperature, °C

0.6

OP

0.6

0.4

70

65 21

20

CO PS

90

System Weight, kg

0.8

System Weight, kg

H. Wang et al. / Energy 36 (2011) 4809e4820

COPS

4816

70

65 5000

Pumping Pressure, kPa Fig. 13. Effect of pump/boiling pressure on overall system COPs and weight.

Fig. 15. Effect of pressure ratio on expander isentropic efficiency.

H. Wang et al. / Energy 36 (2011) 4809e4820

4817

Table 3 Typical state parameters with the 2nd power recuperator. State point (Fig. 17)

7

8

9

10

11

12

15

Pressure (kPa) Temperature (
C)

562.1 95.2

562.1 98.8

562.1 79.6

3200 69.4

3200 94.5

3200 121.6

562.1 101.5

recuperator weight increases with effectiveness as it gets bigger, the system weight declines until the effectiveness is about 70% due to reduction in other power component weights associated with increasing efficiency. The increase in overall weight with effectiveness remains small until about 85% effectiveness is reached. The evaporating and condensing temperatures of 18
C and 67
C were chosen based on design conditions where the indoor air temperature is 32
C and outdoor air temperature is 48.9
C. These temperatures represent a difference from ambient conditions of 14
C for the evaporator and 18.1
for the condenser, which can be reduce to further improve cycle performance. However, unlike subcooling, improved evaporator performance could come at a cost to weight for conventional tube-fin coils, even though the boiler and expander are reduced in size in portion to the efficiency gain. In addition, the amount of air required to move through each device would also need to increase proportionally, which could increase the fan power significantly and thus erode the benefits of lowering the temperature differences. Alternatively, microchannel heat exchangers, as an emerging technology, provide excellent heat transfer characteristics without sacrificing size and weight. They would be excellent choices for the evaporator and condenser. In the meantime, fan power is decreased due to reduced flow passage and improved flow without separation. They have gained more and more attentions across the industries, since it started as an improvement technology for automobile applications when R12 (Freon) was replaced with R134a. It has been widely accepted as microchannel condensers, where size and weight are essential. Both microchannel condensers and evaporators are based on socalled BAM coils, which are constructed of extruded aluminum tubes containing many small refrigerant ports. It greatly enhances the heat transfer characteristics compared to conventional tube-fin

Fig. 16. Schematics 3D Model of BAM Coil Cut-Away View.

of expander is held constant. However, as pumping pressure is increased, it is very difficult to maintain the same expander efficiency with other penalties. Fig. 15 shows the isentropic efficiency of a recently tested scroll expander as a function of PR (pressure ratio), which is defined as the boiling pressure divided by the condensing pressure. It shows that expander isentropic efficiency dropped significantly as pressure ratio or pump exit pressure increased due to under-expansion. As a consequence as shown in Fig. 13, COPs decrease while weight increases with increasing pump or boiling pressure. In order to maintain or achieve higher COPs with increasing pump pressure, multi-stage scroll expander is needed (similar to multi-stage expansions in gas turbines) with more weight. According to the data shown in Fig. 15, the isentropic efficiency of the scroll expander is generally pretty high. The highest isentropic efficiency happens when pressure ratio is around 3.3, which is likely to be the intrinsic expansion ratio of the device. As mentioned earlier, the presence of the power recuperator increases gross COP by 30e40%. As seen in Fig. 14, the COPs increase nearly linearly with increasing effectiveness. Although the

13

Boiler Expander

Qin,hot 14

1st Power Recuperator

12

11 15 8

Compressor 7

2nd Power Recuperator 9

Condenser

Evaporator

5

Cooling Recuperator

10

Qout

6

Pump

Qin,cool Receiver

mp 4

3

Expansion Valve

mc

2

1

Subcooler

Fig. 17. Schematic flow diagram of the combined ORC-VCC cycle with the 2nd power recuperator.

4818

H. Wang et al. / Energy 36 (2011) 4809e4820

Condenser

15

Recuperator

12

Qout Qin,hot

Boiler Expander

11

Receiver

13 14 10

Pump

7

Condenser 6

Evaporator

5

Recuperator

Qout

Compressor

Qin,cool Receiver 4

3

2

1

Expansion Valve Subcooler Fig. 18. Schematic flow diagram of the combined two-fluid ORC-VCC cycle.

coils, thus reducing the thermal resistance between refrigerant and air. A BAM coil can deliver greater cooling capacity than a conventional tube-fin coil for a given space. In other words, a BAM coil can be significantly smaller than a tube-fin coil with the same capacity. In addition, the whole device is made out of aluminum that makes BAM coils significantly lighter than their traditional counterparts. Considerably smaller internal volume for BAM coils provides additional advantages including less refrigerant charge and safer for the environment. A 3D cut-away view of the BAM coil is illustrated in Fig. 16. The air side passages with louver fins are generally 25.4 mm (1 inch) long. 4. Advanced cycle concepts 1) 2nd power recuperator With the cooling cycle recuperation (see Fig. 6), the fluid returns to compressor much warmer than it would with normal superheating. Thus the fluid exits the compressor much warmer due to compression heat, which leads to compressed vapor temperature approaching 95
C. This is well above the condensation temperature, and thus there is significant amount of sensible heat available for preheating the boiler feed stream. As the schematic shows in Fig. 17, the vapor exiting the primary power recuperator in the ORC is combined with the vapor exiting the compressor. The mixture is then fed into a secondary power recuperator. The reason for combining the cooling flow with the power flow in the secondary power recuperator is that it results in a closer match of the heat capacity of the streams on each side of the recuperator. This is because the specific heat of the vapor is lower than the liquid, so raising the mass flow of the vapor brings the overall heat capacity of each stream closer. Reduced heat capacity mismatch provides the potential for reduced heat transfer irreversibility. The typical pressures and temperatures at state points around the 1st and 2nd power recuperators are shown in Table 3. As noticed, pressure drops were

ignored across each component. An alternate configuration would use just the cooling flow in secondary recuperator, which will not have potential issues associated with mixing. 2) Two-fluid system The two-fluid system uses different working fluids in the power generation and cooling portions of the cycle. The flow schematic diagram is illustrated in Fig. 18. This approach has several advantages. It may allow for higher efficiency than a single-fluid cycle of the same weight. This is because the power cycle is less sensitive to the condensing temperature than is the cooling cycle. By keeping the fluids separate, the temperature inside each condenser can be optimized to provide the best system performance and light weight combination. Keeping the fluids separate also allows different lubricants to be used in each section that are tailored to the operating conditions in the expander and compressor (roughly 175
C vs. 75
C). There is also no concern about lubricant accumulating in one device while the other runs dry. Thus lubricant management is easier. Lastly, the single-fluid system requires a low pressure in the evaporator to avoid excessively high pressures in the boiler. Low vapor pressure in the evaporator makes its performance sensitive to pressure drop inside the device. This requirement makes the evaporator design for the single-fluid system significantly more challenging. On the other hand, two-fluid system also has its disadvantages. For example, shaft seals are required to isolate the expansion and compression sections; more fluids need to deal with, which could be detraction from a service perspective. 5. Conclusions The study introduces a novel thermally activated cooling concept e a combined cycle couples an ORC and a VCC (vapor compression cycle). The system can be powered by solar thermal, geothermal or various waste heat streams. A systematic design

H. Wang et al. / Energy 36 (2011) 4809e4820

study was conducted to investigate effects of a variety of cycle configurations on system performance in terms of cycle gross COP. Compared to the basic cycle, an advanced cycle with subcooling and cooling recuperation achieves 22% improvement on gross COP at high indoor and outdoor temperatures. A parametric trade-off study was conducted for the advanced cycle, in order to find the critical design parameters affecting cycle COPs and system weight the most. While increasing the expander isentropic efficiency shows the most positive outcomes, the effects of changing condensing temperature, evaporating temperature, pump/boiling pressure and effectiveness of the power recuperator have mixed results. Two additional cycle concepts with either potentially higher COP or practical advantages were also proposed, including one concept of adding a secondary recuperator in the ORC and the other concept of using different working fluids in the power and cooling cycles, or so-called dual-fluid system. References [1] Prigmore D, Barber R. Cooling with the sun’s heat design considerations and test data for a Rankine cycle prototype. Solar Energy 1975;17:185e92. [2] Barber R. Current costs of solar powered organic Rankine cycle engines. Solar Energy 1978;20:1e6. [3] Kane M. Small hybrid solar power system. Energy 2003;28:1427e43. [4] Delgado-Torres AM, García-Rodríguez L. Analysis and optimization of the lowtemperature solar organic Rankine cycle (ORC). Energy Conversion and Management 2010;51:2846e56. [5] Pei G, Li J, Ji J. Analysis of low temperature solar thermal electric generation using regenerative organic Rankine cycle. Applied Thermal Engineering 2010; 30:998e1004. [6] Tchanche BF, Papadakis G, Lambrinos G, Frangoudakis A. Fluid selection for a low-temperature solar organic Rankine cycle. Applied Thermal Engineering 2009;29:2468e76. [7] Jing L, Gang P, Jie J. Optimization of low temperature solar thermal electric generation with organic Rankine cycle in different areas. Applied Energy 2010;87:3355e65. [8] Palosojr G, Mohanty B. Cascading vapour absorption cycle with organic Rankine cycle for enhancing geothermal power generation. Renewable Energy 1993;3:669e81. [9] Mohanty B, Palosojr G. Economic power generation from low-temperature geothermal resources using organic Rankine cycle combined with vapour absorption chiller. Heat Recovery Systems and CHP 1992;12:143e58. [10] Madhawahettiarachchi H, Golubovic M, Worek W, Ikegami Y. Optimum design criteria for an organic Rankine cycle using low-temperature geothermal heat sources. Energy 2007;32:1698e706. [11] Heberle F, Brüggemann D. Exergy based fluid selection for a geothermal organic Rankine cycle for combined heat and power generation. Applied Thermal Engineering 2010;30:1326e32. [12] Astolfi M, Xodo L, Romano MC, Macchi E. Technical and economical analysis of a solaregeothermal hybrid plant based on an organic Rankine cycle. Geothermics; 2010. [13] Srinivasan KK, Mago PJ, Krishnan SR. Analysis of exhaust waste heat recovery from a dual fuel low temperature combustion engine using an organic Rankine cycle. Energy 2010;35:2387e99. [14] Wei D, Lu X, Lu Z, Gu J. Dynamic modeling and simulation of an organic Rankine cycle (ORC) system for waste heat recovery. Applied Thermal Engineering 2008;28:1216e24. [15] Larjola J. Electricity from industrial waste heat using high-speed organic Rankine cycle (ORC). International Journal of Production Economics 1995;41: 227e35. [16] Quoilin S, Lemort V, Lebrun J. Experimental study and modeling of an organic Rankine cycle using scroll expander. Applied Energy 2010;87:1260e8. [17] Dai Y, Wang J, Gao L. Parametric optimization and comparative study of organic Rankine cycle (ORC) for low grade waste heat recovery. Energy Conversion and Management 2009;50:576e82. [18] Roy J, Mishra M, Misra A. Parametric optimization and performance analysis of a waste heat recovery system using organic Rankine cycle. Energy; 2010. [19] Wei D, Lu X, Lu Z, Gu J. Performance analysis and optimization of organic Rankine cycle (ORC) for waste heat recovery. Energy Conversion and Management 2007;48:1113e9. [20] Hung T. Waste heat recovery of organic Rankine cycle using dry fluids. Energy Conversion and Management 2001;42:539e53. [21] Wang H, Peterson RB, Herron T. Experimental performance of a compliant scroll expander for an organic Rankine cycle. Proceedings of the Institution of Mechanical Engineers Part A: Journal of Power and Energy 2009;223:863e72. [22] Peterson RB, Wang H, Herron T. Performance of a small-scale regenerative Rankine power cycle employing a scroll expander. Proceedings of the Institution of Mechanical Engineers Part A: Journal of Power and Energy 2008;222: 271e82.

4819

[23] Desai NB, Bandyopadhyay S. Process integration of organic Rankine cycle. Energy Oct. 2009;34:1674e86. [24] Vaja I, Gambarotta A. Internal combustion engine (ICE) bottoming with organic Rankine cycles (ORCs). Energy Feb. 2010;35:1084e93. [25] Schuster A, Karellas S, Aumann R. Efficiency optimization potential in supercritical organic Rankine cycles. Energy Feb. 2010;35:1033e9. [26] Wang H, Peterson R, Harada K, Miller E, Ingram-Goble R, Fisher L, et al. Performance of a combined organic Rankine cycle and vapor compression cycle for heat activated cooling. Energy Jan. 2011;36:447e58. [27] Manolakos D, Papadakis G, Mohamed E, Kyritsis S, Bouzianas K. Design of an autonomous low-temperature solar Rankine cycle system for reverse osmosis desalination. Desalination 2005;183:73e80. [28] Manolakos D, Papadakis G, Kyritsis S, Bouzianas K. Experimental evaluation of an autonomous low-temperature solar Rankine cycle system for reverse osmosis desalination. Desalination 2007;203:366e74. [29] Garciarodriguez L, Blancogalvez J. Solar-heated Rankine cycles for water and electricity production: POWERSOL project. Desalination 2007;212:311e8. [30] Delgadotorres A, Garciarodriguez L. Double cascade organic Rankine cycle for solar-driven reverse osmosis desalination. Desalination 2007;216:306e13. [31] Delgadotorres A, Garciarodriguez L, Romeroternero V. Preliminary design of a solar thermal-powered seawater reverse osmosis system. Desalination 2007;216:292e305. [32] Bruno J, Lopezvillada J, Letelier E, Romera S, Coronas A. Modelling and optimisation of solar organic Rankine cycle engines for reverse osmosis desalination. Applied Thermal Engineering 2008;28:2212e26. [33] Kosmadakis G, Manolakos D, Kyritsis S, Papadakis G. Economic assessment of a two-stage solar organic Rankine cycle for reverse osmosis desalination. Renewable Energy 2009;34:1579e86. [34] Manolakos D, Kosmadakis G, Kyritsis S, Papadakis G. Identification of behaviour and evaluation of performance of small scale, low-temperature organic Rankine cycle system coupled with a RO desalination unit. Energy 2009;34:767e74. [35] Tchanche B, Lambrinos G, Frangoudakis A, Papadakis G. Exergy analysis of micro-organic Rankine power cycles for a small scale solar driven reverse osmosis desalination system. Applied Energy 2010;87:1295e306. [36] Kosmadakis G, Manolakos D, Papadakis G. Parametric theoretical study of a two-stage solar organic Rankine cycle for RO desalination. Renewable Energy 2010;35:989e96. [37] Nafey A, Sharaf M, García-Rodríguez L. Thermo-economic analysis of a combined solar organic Rankine cycle-reverse osmosis desalination process with different energy recovery configurations. Desalination 2010;261:138e47. [38] Nafey A, Sharaf M. Combined solar organic Rankine cycle with reverse osmosis desalination process: energy, exergy, and cost evaluations. Renewable Energy 2010;35:2571e80. [39] Delgado-Torres AM, García-Rodríguez L. Preliminary design of seawater and brackish water reverse osmosis desalination systems driven by lowtemperature solar organic Rankine cycles (ORC). Energy Conversion and Management 2010;51:2913e20. [40] Herold K, Howe L, Radermacher R. Analysis of a hybrid compressiondabsorption cycle using lithium bromide and water as the working fluid. International Journal of Refrigeration 1991;14:264e72. [41] Medrano M. Double-lift absorption refrigeration cycles driven by lowetemperature heat sources using organic fluid mixtures as working pairs. Applied Energy 2001;68:173e85. [42] Ezzine N, Barhoumi M, Mejbri K, Chemkhi S, Bellagi A. Solar cooling with the absorption principle: first and second law analysis of an ammoniadwater double-generator absorption chiller. Desalination 2004;68:137e44. [43] Mroz T. Thermodynamic and economic performance of the LiBreH2O single stage absorption water chiller. Applied Thermal Engineering 2006;26: 2103e9. [44] Kim B, Park J. Dynamic simulation of a single-effect ammoniaewater absorption chiller. International Journal of Refrigeration 2007;30:535e45. [45] Kim D, Infante Ferreira C. Air-cooled LiBrewater absorption chillers for solar air conditioning in extremely hot weathers. Energy Conversion and Management 2009;50:1018e25. [46] Jiangzhou S. Experimental research on characteristics of corrosion-resisting nickel alloy tube used in triple-effect LiBr/H2O absorption chiller. Applied Thermal Engineering 2001;21:1161e73. [47] Kim J. Simulation of the compressor-assisted triple-effect H2O/LiBr absorption cooling cycles. Applied Thermal Engineering 2002;22:295e308. [48] Manohar H, Saravanan R, Renganarayanan S. Modelling of steam fired double effect vapour absorption chiller using neural network. Energy Conversion and Management 2006;47:2202e10. [49] Figueredo G, Bourouis M, Coronas A. Thermodynamic modelling of a twostage absorption chiller driven at two-temperature levels. Applied Thermal Engineering 2008;28:211e7. [50] Torrella E, Sánchez D, Cabello R, Larumbe J, Llopis R. On-site real-time evaluation of an air-conditioning direct-fired double-effect absorption chiller. Applied Energy 2009;86:968e75. [51] Wang J, Zheng D. Performance of one and a half-effect absorption cooling cycle of H2O/LiBr system. Energy Conversion and Management 2009;50: 3087e95. [52] Garimella S. Performance evaluation of a generator-absorber heat-exchange heat pump. Applied Thermal Engineering 1996;16:591e604. [53] Kang Y. An advanced GAX cycle for waste heat recovery: WGAX cycle. Applied Thermal Engineering 1999;19:933e47.

4820

H. Wang et al. / Energy 36 (2011) 4809e4820

[54] Velázquez N. Methodology for the energy analysis of an air cooled GAX absorption heat pump operated by natural gas and solar energy. Applied Thermal Engineering 2002;22:1089e103. [55] Sabir H. Analytical study of a novel GAX-R heat driven refrigeration cycle. Applied Thermal Engineering 2004;24:2083e99. [56] Rameshkumar A, Udayakumar M. Simulation studies on GAX absorption compression cooler. Energy Conversion and Management 2007;48:2604e10. [57] Gomez V, Vidal A, Best R, Garciavalladares O, Velazquez N. Theoretical and experimental evaluation of an indirect-fired GAX cycle cooling system. Applied Thermal Engineering 2008;28:975e87. [58] Park C, Koo J, Kang Y. Performance analysis of ammonia absorption GAX cycle for combined cooling and hot water supply modes. International Journal of Refrigeration 2008;31:727e33. [59] Jawahar C, Saravanan R. Generator absorber heat exchange based absorption cycleda review. Renewable and Sustainable Energy Reviews 2010;14:2372e82. [60] Fong K, Chow T, Lee C, Lin Z, Chan L. Advancement of solar desiccant cooling system for building use in subtropical Hong Kong. Energy and Buildings 2010; 42:2386e99. [61] Dai Y. Use of liquid desiccant cooling to improve the performance of vapor compression air conditioning. Applied Thermal Engineering 2001;21:1185e202. [62] Mavroudaki P, Beggs C, Sleigh P, Halliday S. The potential for solar powered single-stage desiccant cooling in southern Europe. Applied Thermal Engineering 2002;22:1129e40. [63] Halliday S. The use of solar desiccant cooling in the UK: a feasibility study. Applied Thermal Engineering 2002;22:1327e38. [64] Kanoglu M. Energy and exergy analyses of an experimental open-cycle desiccant cooling system. Applied Thermal Engineering 2004;24:919e32. [65] Carpinlioglu M, Yildirim M. A methodology for the performance evaluation of an experimental desiccant cooling system. International Communications in Heat and Mass Transfer 2005;32:1400e10. [66] Yin Y, Zhang X, Chen Z. Experimental study on dehumidifier and regenerator of liquid desiccant cooling air conditioning system. Building and Environment 2007;42:2505e11. [67] Bourdoukan P, Wurtz E, Joubert P. Experimental investigation of a solar desiccant cooling installation. Solar Energy 2009;83:2059e73. [68] La D, Dai Y, Li Y, Wang R, Ge T. Technical development of rotary desiccant dehumidification and air conditioning: a review. Renewable and Sustainable Energy Reviews 2010;14:130e47.

[69] Ge T, Ziegler F, Wang R, Wang H. Performance comparison between a solar driven rotary desiccant cooling system and conventional vapor compression system (performance study of desiccant cooling). Applied Thermal Engineering 2010;30:724e31. [70] Panaras G, Mathioulakis E, Belessiotis V, Kyriakis N. Theoretical and experimental investigation of the performance of a desiccant air-conditioning system. Renewable Energy 2010;35:1368e75. an E, Büyükalaca O, Yılmaz T, Hepbas¸lı A. Experimental investigation [71] Hürdog of a novel desiccant cooling system. Energy and Buildings 2010;42: 2049e60. [72] La D, Dai Y, Li Y, Ge T, Wang R. Study on a novel thermally driven air conditioning system with desiccant dehumidification and regenerative evaporative cooling. Building and Environment 2010;45:2473e84. [73] Eicker U, Schneider D, Schumacher J, Ge T, Dai Y. Operational experiences with solar air collector driven desiccant cooling systems. Applied Energy 2010;87: 3735e47. [74] Heidarinejad G, Pasdarshahri H. The effects of operational conditions of the desiccant wheel on the performance of desiccant cooling cycles. Energy and Buildings 2010;42:2416e23. [75] Cui Q. Performance study of new adsorbent for solid desiccant cooling. Energy 2005;30:273e9. [76] Jia C, Dai Y, Wu J, Wang R. Use of compound desiccant to develop high performance desiccant cooling system. International Journal of Refrigeration 2007;30:345e53. [77] Tretiak C, Abdallah NB. Sorption and desorption characteristics of a packed bed of clayeCaCl2 desiccant particles. Solar Energy 2009;83:1861e70. [78] Jia C, Dai Y, Wu J, Wang R. Experimental comparison of two honeycombed desiccant wheels fabricated with silica gel and composite desiccant material. Energy Conversion and Management 2006;47:2523e34. [79] Ge T, Li Y, Dai Y, Wang R. Performance investigation on a novel two-stage solar driven rotary desiccant cooling system using composite desiccant materials. Solar Energy 2010;84:157e9. [80] Ramzy AK, Kadoli R, Ashok Babu T. Improved utilization of desiccant material in packed bed dehumidifier using composite particles. Renewable Energy 2011;36:732e42. [81] Aphornratana S, Sriveerakul T. Analysis of a combined Rankineevapourecompression refrigeration cycle. Energy Conversion and Management 2010;51: 2557e64.