Energy, exergy and environmental analysis of a hybrid combined cooling heating and power system utilizing biomass and solar energy

Energy Conversion and Management 124 (2016) 566–577

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Energy Conversion and Management
journal homepage: www.elsevier.com/locate/enconman

Energy, exergy and environmental analysis of a hybrid combined cooling
heating and power system utilizing biomass and solar energy
Jiangjiang Wang ⇑, Ying Yang
School of Energy, Power and Mechanical Engineering, North China Electric Power University, Baoding, Hebei Province 071003, China

a r t i c l e

i n f o

Article history:
Received 23 May 2016
Received in revised form 8 July 2016
Accepted 21 July 2016

Keywords:
Combined cooling heating and power
(CCHP) system
Biomass energy
Solar energy
Thermodynamics analysis
Energy complementarity

a b s t r a c t
A hybrid combined cooling heating and power (CCHP) system driven by biomass and solar energy is

proposed, and their complementarity to enhance the system’s energy efficiency is analyzed and shown.
The CCHP system is primarily composed of a biomass gasification sub-system, solar evacuated collector,
internal combustion engine and dual-source powered mixed-effect absorption chiller. The product gas
produced by the gasifier drives the internal combustion engine to generate power, and the waste heat
after generation is utilized to produce cooling and heating with the collected heat from the solar collectors. Under the design conditions, the thermodynamic performances under variable external conditions
and energy ratios are investigated and analyzed. The results indicate that the primary energy ratio and
the exergy efficiency are 57.9% and 16.1%, respectively, and the carbon emission reduction ratio is about
95.7%, at the design condition. The complementarity analysis between the biomass and solar energy
shows that the biomass subsystem makes a greater contribution to the total system primary energy ratio
and exergy efficiency than the contributions from the solar subsystem, and the participation of solar
energy is conducive to the system emission reduction.
Ó 2016 Elsevier Ltd. All rights reserved.

1. Introduction
Distributed energy systems (DES) are becoming one of the more
attractive options worldwide because of their high overall efficiency, low greenhouse gas emissions, high reliability and other
features [1]. A DES, which includes combined heating and power
(CHP) system, combined cooling, heating and power (CCHP) system, and distributed renewable energy technologies can realize a
cascading utilization of energy. The advance and development of
DES has promoted various studies on their technology [2], system
configuration [3,4], performance evaluation [5], and optimization
[6,7], and most of the studies have concentrated on establishing
optimal DES to achieve favorable costs, energy savings and emission reductions.
In particular, renewable energy resources are sustainable alternatives to natural gas for driving traditional CHP/CCHP systems [8],
which has gradually become a topic of intense study. Focusing on
the energy sources in DES, hybrid DES combine renewable energy
resources and fossil resources to decrease greenhouse gas emissions and simultaneously accommodate instabilities in renewable
energy. The literature on hybrid DES discusses different forms of

⇑ Corresponding author.

E-mail address: (J. Wang).
/>0196-8904/Ó 2016 Elsevier Ltd. All rights reserved.

complementary energy, for example hybrid wind/photovoltaic
energy systems [9], multicomponent systems, including photovoltaic panels, wind generators and biomass gasification plants
[10], hybrid geothermal-solar systems [11], hybrid solar and chemical looping combustion systems [12], CCHP systems based on cofiring natural and biomass gasification gases [13], solar-biomass
hybrid air-conditioning systems [14] and hybrid polygeneration
systems that utilize biomass fuel and solar power [15].
Among the renewable energy resources, biomass and solar
energy currently have attracted considerable attention from academics and researchers for their green environmental protection
and inexhaustibility advantages. Moreover, biomass is a stable
energy resource that can produce continuous power and simultaneously reduce carbon dioxide (CO2) emissions. To date, only a
few studies have been conducted to explore hybrid system driven
by biomass and solar energy, especially in terms of analyzing the
hybrid proportion of energy resources and showing how they can
provide complementary sources of energy. Wang et al. [13] analyzed the influence of different mixture ratios of natural gas and
biogas on thermodynamic performance and exergoeconomic cost
and discussed the complementary performances of biomass and
natural gas. Hashim et al. [16] used the concept of an IBS (Integrated Biomass Solar) town and developed a hybrid solar and biomass plant, which, however, focused on the complementarity of


J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

567

Nomenclature
CCHP
CHP
COP
CO2

DES
HHV
HX
IBS
ICE
LHV
RMSE

combined cooling heating and power
combined heating and power
coefficient of performance
carbon dioxide
distributed energy system
higher heating value
heat exchanger
integrated biomass solar
internal combustion engine
lower heating value
root mean square error

Symbols
A
E
EX
HHV
LHV
_
m
P
Q

T

area (m2)
electricity (kW)
exergy (kW)
higher heating value (MJ kgÀ1 or MJ NmÀ3)
lower heating value (MJ kgÀ1 or MJ NmÀ3)
mass flow rate (kg/s)
pressure (kPa)
energy(kW)
temperature (K)

the electrical supply rather than a system thermodynamics analysis. Academics and specialists have conducted a variety of relevant
solar and biomass energy research, including the optimal design of
a hybrid solar-assisted biomass energy system for heating [17], a
hybrid solar and biomass energy complementary system for power
generation [18], a hybrid solar and chemical looping combustion
system for solar thermal energy storage [12] and a study of a
hybrid solar-biomass air-conditioning system for cooling [14].
However, that research has primarily focused only on particular
energy supply products and rarely concentrated on the combined
supply of cooling, heating and power. Based on those considerations, the present study is motivated to explore this issue.
The main aim of this work is to propose a hybrid CCHP system
that is driven by biomass and solar energy and to explore the complementarity of biomass and solar energy on the energy efficiency.
Four complementary conditions are discussed, variable solar irradiation, variable power loads, variable biomass input and variable
solar energy input. The first two conditions are studied under single variable, respectively, and the subsequent two variables are
conducted simultaneously to analyze all combinations of biomass
and solar energy input. In addition, to evaluate the specific influences of biomass and solar subsystems, we propose the concept
of subsystem contribution in energy efficiency. Therefore, thermodynamic models of a hybrid CCHP system were constructed and
validated; those models used existing technologies of solar heat

collection, biomass gasification, absorption refrigeration and
power generation. Performances under varying operating conditions were then analyzed, and the system thermodynamic performance, including the primary energy ratios and exergy
efficiencies under different energy proportions are discussed to
determine the energy efficiency enhancement mechanism in the
hybrid CCHP system. The hybrid CCHP system offers several advantageous features, including (1) combined two kinds of renewable
energy which was environmentally friendly, (2) reduced the consumption of fossil energy, (3) revealed the complementarity of biomass and solar energy that benefit for the system optimization.
This hybrid CCHP system can be innovatory in combined application of biomass and solar energy, especially suitable for remote
areas where there are sufficient crops and solar energy.

z

g

mass fraction (dimensionless)
efficiency

Subscripts
b
biomass
bio
biomass
c
cooling
ch
chemical
e
electricity
ee
energy
ex

exergy
exh
exhaust
f
fuel
h
heating
hw
hot water
jw
jacket water
n
nominal
p
pump
s
solar
sol
solar
w
water
rw
refrigeration water
v
variable

2. System description
The flowchart of a hybrid CCHP system driven by biomass and
solar energy is shown in Fig. 1; the system is composed of a biomass gasification subsystem, solar photothermal collection subsystem, internal combustion engine (ICE) power subsystem and waste
heat utilization subsystem. Biomass material is first gasified in the

downdraft gasifier, and then its product gas is sent to be further
cooled in cyclone and purified in spray scrubbers, respectively.
Subsequently, the product gas fuel drives the ICE to generate
power. During this process, the heat exchanger (HX-01) is
employed to recover the sensible heat from the product gas exiting
the gasifier to produce domestic hot water. The solar evacuated
collectors are used to collect solar photothermal energy to produce
mesothermal hot water, the outlet temperature of which is
designed to match the outlet temperature of the jacket water from
the ICE at approximately 85 °C.
The mixture of jacket water and solar hot water cooperates with
the exhaust gas from the ICE, which has a temperature of approximately 460 °C, is fed to a dual-source powered mixed-effect LiBrH2O absorption chiller to produce chilled water. After releasing heat
in absorption chiller, the outlet exhaust gas still has a temperature
of approximately 170 °C, and the heat exchanger (HX-02) is therefore used to recover the waste heat to preheat the cool water.
Regarding the hot water part, the outlet temperature, which is
approximately 70 °C, is split into two streams that return to the collector and jacket, respectively, for the next cycle. Moreover, when
the temperature of the cooled jacket water cannot meet the requirement on engine cooling, it can be further cooled in cooling tower 01.
Consequently, the system creates three products: electricity,
chilled water and hot water. Furthermore, the absorption chiller
can be used as a heat exchanger to produce hot water, and two
products, electricity and hot water, are generated. For later analysis, the base design parameters are shown in Table 1.
3. Thermodynamic model
The thermodynamic models (biomass gasification, ICE, solar
evacuated collectors, dual-source absorption chiller and heat


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J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

Fig. 1. Schematic of a hybrid CCHP system driven by biomass and solar energy.

Table 1
Base design parameters.
Parameter

Value

HX-01 outlet temperature (°C)
Chilled water temperature (°C)
Hot water temperature (°C)
Cooling water temperature (°C)
Jacket water temperature (°C)
Cool water temperature (°C)
Chiller exhaust gas temperature
HX-02 exhaust gas temperature
Efficiency of gas-water heat exchanger

200 (state 5)
7/14 (states 22/21)
60 (state 31)
32/36 (states 27/28)
70/85(states 13/12)
25 (state 29)
170 (state 19)
120 (state 20)
0.90

exchanger) are presented in this section. The biomass gasification,
ICE and heat exchanger were constructed following [13], and they

are briefly introduced.
3.1. Biomass gasification
In the literature, there are several studies on various models of
biomass gasification [19–22]. The thermochemical equilibrium
model for biomass air gasification in [13] contains pyrolysis and
gasification modules and considers the residual tar and char, which
were simulated and validated for biomass air gasification

preferably. Using that biomass gasification model, wood chips as
a biomass material (Table 2) are gasified to produce the product
gas in Table 2. In which, the characteristics of wood chips and
the gasification process are assumed to be constant, and the property of the product gas is considered stable.
The exergy of biomass can be calculated as [23]:

EX biomass ¼ mbiomass

ð1 À zMoisture À zAsh Þ Â bLHV biomass
þzMoisture  exch;water þ zAsh  exch;Ash

1:044 þ 0:0160ðzH =zC Þ À 0:3493ðzO =zC Þ
½1 þ 0:0531ðzH =zC ފ þ 0:0493ðzN =zC Þ
b¼
1 À 0:4124ðzO =zC Þ

!
and

ð1Þ

!

ð2Þ

where mbiomass is the mass flow rate of biomass, zMoisture, zAsh, zH, zC,
zO and zN are the mass fraction of moisture, ash, hydrogen, carbon,
oxygen and nitrogen of biomass, respectively. ech,water and ech,Ash are
the chemical exergy of water and ash, in this paper, they are 1300
and 0 kJ/kmol, respectively.
3.2. Internal combustion engine
An appropriate model for an ICE will provide realistic estimates
of performance/efficiency maps for both electrical power output


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J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577
Table 2
Properties of biomass material and product gas [13].
Fuel

Parameters

Wood chips

Proximate analysis (wt%)

Volatile
70.2

Fixed carbon
28.3

Ash (zAsh)
1.5

Moisture (zMoisture)
15

Ultimate analysis (wt%)

C (zC)
51.2

H (zH)
6.1

O (zO)
40.9

N (zN)
0.3

HHV (MJ/kg)

20.43

LHV (MJ/kg)

19.09

Composition analysis (mol%)

CO
12.05

H2
12.32

CO2
15.21

LHV (MJ/Nm3)

3.12

Product gas

and useful thermal output for various capacities of engines. Different modified methods have been proposed to predict engine performance [24,25]. Furthermore, the characteristics of an ICE
driven by fuel with lower heating value are dramatically different
from those of a traditional ICE model [26]. By modeling the error
between the actual fuel and design fuel, Wang et al. developed
an ICE model driven by the product gas instead of natural gas
[13] and is the model adopted in this paper.
For ICE which driven by lower heating value fuel, the generation
efficiency (ge) can be calculated as [24,27]:



ge ¼ 0:102



LHV f
0:0563
þ 0:897 Â 28:08ðNÃe Þ
LHV NG

ð3Þ

where LHVf and LHVNG are the lower heating value of product gas
and natural gas, respectively. Ne⁄ is the nominal generation capacity
of ICE, which is 1.1 times of practical generating volume.
Similarly, the outlet temperature of exhaust gas from the ICE
can be expressed as [24]:

T ex



LHV f
¼ 0:025
þ 0:974
LHV NG
h
i
2
 2  10À5 ðNÃe Þ À 0:0707NÃe þ 758:33

ð4Þ

In terms of the recovery efficiency of exhaust gas (gr, exh) and
jacket water (gr, jw), assume the practical recuperated heat equals

Heat pipe condenser

CH4
1.1

to the nominal recuperated heat, the gr,
lated as:

gr;exh

N2
59.33

exh

À
Á
Q exh mexh hexh;i À hexh;o
¼
¼
Qf
Qf

gr;jw ¼

À
Á
Q jw mjw cp;jw T jw;o À T jw;i
¼

Qf
Qf

and gr,

jw

can be calcu-

ð5Þ

ð6Þ

where Qf is the energy of input fuel, mexh and mjw are the mass flow
rate of exhaust gas and jacket water, respectively. hexh,i and hexh,o are
the empathy of inlet and outlet exhaust gas of absorption chiller,
respectively. cp,jw is the average specific heat of jacket water and
the Tjw,o and Tjw,i are the outlet and inlet temperature of cooling
water of jacket, respectively.
3.3. Solar evacuated tube collector
Among the various forms of solar collectors on the market, the
advantages of an evacuated tube collector lie in its simple installation and high operating temperature and thermal efficiency, especially given the wide adaptability in solar irradiance, and is more
appropriate for DES in various types of regions. The detailed structure of the solar evacuated tube collector is shown in Fig. 2 [28]. The
collector features a heat pipe (a highly efficient thermal conductor)
placed inside a vacuum-sealed tube (evacuated tube). The heat

Manifold
Fluid flow

Evacuated tube

Absorber plate
Heat pipe evaporator
Cross-sectional detail

Fig. 2. Diagram of the evacuated tube collector [29].


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J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

pipe, which is a sealed copper pipe, is then attached to a black copper fin that fills the tube (absorber plate). The liquid–vapor phase
change materials (water) that are used to transfer heat undergo
an evaporating–condensing cycle in the heat pipe. In that cycle,
the liquid is evaporated by the solar heat, and the vapor then rises
to the heat sink region, condensing and releasing its latent heat.
Subsequently, the condensed fluid falls back to the bottom of the
heat pipe, and the process repeats. In the heat exchange process,
the metal tip projects into a heat exchanger (manifold), and, when
the working fluid (water) flows through the manifold, it will pick up
heat from the tubes and gives off its heat in the next procedure.
3.3.1. Heat transfer analysis
The instantaneous efficiency of the evacuated tube collector can
be expressed as [28]:

g¼

!
AP
Ti À Ta

;
F R ðsaÞe À U L
As
G

where AP and As are the absorber plate area and insolation area,
respectively (m2). (sa)e is the efficient fraction of the incident solar
energy ultimately absorbed by the absorber plate (s is the transmissivity, and a is the absorptivity), and Ti and Ta are the inlet water
flow and ambient temperatures, respectively (K). G is the irradiance
of the total solar radiation on the horizontal surface. FR is the heat
removal factor which is concerned with the collector structure,
the detail formulas can be found in [30]. UL is the overall heat loss
coefficient (W/(m2ÁK)), and the calculation method is as follows.
The thermal analysis of collector is based on heat transfer theory, which considers the convective heat transfer between the collector and ambient fluid, radiative transfer between the absorber
plate and the ambient fluid and the heat transfer of the twophase flow in the slender heat pipe. In order to simplify the heat
transfer of evacuated tube collector, generally the basically
assumptions are adopted as follows:
(1) Ignore the convective heat transfer and heat conduction
between the rare air and tube wall in evacuated tube.
(2) Ignore the contact thermal resistance between the absorber
plate and the evaporation section of heat pipe, and between
the manifold wall and condensation section of heat pipe.
(3) Assume the heat transfer from heat pipe to the ambient
keeps exclusively radial transfer way.
(4) Assume the water in heat pipe, ambient air around the tube
wall are both under steady flow.
(5) Ignore the micro heat transfer with regard to wick conduction or other wispy liquid form, include but not limits to
vapor-liquid and liquid-vapor interface thermal resistance,
coefficient of internal thermal resistance into the evaporator.
To distinctly illustrate the heat transfer relationship, a thermal

network is given, as is shown in Fig. 3. Where Tp, Tg is the temperature of absorber plate and glass tube, respectively (K); Ub is the
convective heat transfer coefficient between the manifold insulating layer and ambient, hp–g is the radiation heat transfer coefficient
between the absorber plate and glass tube inside wall, and hr,g–a
and hc,g–a are radiation heat transfer coefficient and convective
heat transfer coefficient between the glass tube outer wall and
ambient, respectively. Qu and QL are useful energy can obtain and
heat losses of the collector, respectively.
Then, the overall heat loss coefficient can be calculated by the
following equations set [28,30]:

UL ¼ Ut þ Ub
Ut ¼

1
hpÀg

þ

ð8Þ
1
hgÀa

Fig. 3. The thermal network of evacuated tube collector.

ð7Þ

!À1
ð9Þ

Ab

Ap

ð10Þ



2rðT p þ T g Þ T 2p þ T 2g


¼
2Ap 1
1
ep þ Ag eg À 1

ð11Þ

Ub ¼

hpÀg

tb
kb

1
þh

Á

1
c;bÀa

hgÀa ¼ hr;gÀa þ hc;gÀa

ð12Þ



hr;gÀa ¼ eg rðT g þ T a Þ T 2g þ T 2a

ð13Þ

hc;gÀa ¼

Nu Á ka
Dg

ð14Þ

Ag hgÀa ðT p À T a Þ ¼ Ap hpÀg ðT P À T g Þ

ð15Þ

where Ut is the gross heat transfer coefficient between tube and
ambient, tb, kb, Ab are thickness, heat conduction coefficient, and
area of manifold insulating layer. Generally choose fiberglass as
insulation material, the thickness is about 0.05 m and the heat conduction coefficient is about 0.048 W/(mÁK) at normal temperature
[30]. hc,b–a is the convective heat transfer coefficient between the
manifold insulating layer and ambient, the typical value range is
1.5–2.0 W/(m2ÁK), thus this paper choose the average value 1.75.
In addition, to ensure the accuracy of the analysis under

dynamic conditions, the heat transfer coefficients of the evaporation section and condensation section in the heat pipe can be
determined [31]:

!
0:23 
0:4
0:3
0:2 
q0:65
kl c0:7
Psat
q
l
pl g
he ¼ 0:32
and
0:4 0:1
Pa
pDe le
q0:25
v hlv ll
À

hc ¼ 0:943

Á

ql ql À qv gk3l hlv
ll ðT v À T c ÞLc

ð16Þ

!0:25
ð17Þ

where ql, kl, ll and cpl are the density (kg/m3), conductivity (W/
(mÁK)), viscosity (kg/(mÁs)) and specific heat (kJ/(kgÁK)), respectively, of liquid water. g (m/s2) is the local acceleration of gravity,
q (W/m2) is the heat flux of the evaporation section, and qv and
Tv are the density and temperature of the vapor, respectively. hlv
(kJ/kg) is the latent heat of vaporization, and Psat and Pa (kPa) are
the pressures of the saturated vapor and ambient, respectively. Tc
and Lc are the temperature and length of the condensation section,
respectively.


J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

3.3.2. Simulation and validation
The detailed structural parameter values of the collector are
shown in Table 3.
Based on these initialization parameters and certain external
conditions, the simulation can be performed using the EES software package. And the performances in each transient point can
be fitted in a linear curve. Due to the restriction of experimental
platform in this research group, this paper identifies the simulation
through the experiment which operated by the solar energy
research institute of Beijing. The experimental data and external
environmental statics can be found in Ref. [30]. A comparison of
the experimental results and the simulated linear curve of the
instantaneous efficiency is shown in Fig. 4. The instantaneous efficiency of the collector is the ratio of useful heat absorbed by the
work fluid to the solar energy which project on the lighting surfaces under a certain approximated steady-state condition, as is

shown in formula (5). The comparison indicates that the trend of
the simulated linear curve is consistent with the distribution of
the experimental data. In addition, the root mean square error
(RMSE) is less than 5%, which confirms the accuracy of the collector
model.
3.4. Absorption chiller
Because there are two streams of waste heat from the ICE, the
exhaust gas and jacket water, the dual-source powered mixedeffect LiBr-H2O chiller shown in Fig. 5 was adopted [29]. The primary parts of the dual-source powered absorption chiller include
a high pressure generator (HG), two low pressure generators
(LG1 and LG2a condenser, throttle, evaporator, absorber, low temperature exchanger (LX) and a high temperature exchanger (HX).
The dual-sources (exhaust gas and hot water) are sent into HG
and LG1, respectively, to evaporate the liquid refrigerant. The

571

refrigerant vapor from HG then flows into the LG2, releasing its
partial condensing heat to generate the third part of refrigerant
vapor. The all-refrigerant vapor from HG, LG1 and LG2 finally flows
into the condenser, where it is cooled and condensed by the cooling water. After a throttling process, the low-temperature and lowpressure liquid refrigerant fall into the evaporator. In the evaporator, by absorbing heat from the chilled water, the liquid refrigerant
evaporates and produces a continuous cooling output. The new
refrigerant vapor will be absorbed by the strong solution in the
absorber and is then pumped into HG and LG1 for the next cycle.
During the desorption and absorption, HX and LX recover heat by
transferring it from the high temperature strong solution to the
low temperature weak solution.
In terms that there are complex influence factors would impact
on the performance, there proposes some commonly adopted
assumptions about the refrigeration process:
(1) Ignore the heat losses in each components and pressure
losses between each connection lines; assume the pressure

difference between the absorption and evaporation is
0.05 kPa and the pressure of condenser is equal to that of
low generator.
(2) The simulation and analysis are both under steady state, the
LiBr-H2O solution are in steady state during the cycle.
(3) The outlet refrigerant steam of evaporator, outlet refrigerant
liquid of condenser, the outlet weak solution of absorber and
the outlet solution of high generator and low generator are
all statured.
(4) Ignore the power consumption of solution pump.
(5) All the heat transfer unit are regarded as countercurrent
heat exchange mode, and use the logarithmic mean temperature difference in heat-transfer calculation.
During the cycle of LiBr-H2O solution, there are continuous
energy and matter enter and leave in each components. The energy
balance, mass balance and solute equilibrium can be summarized
as [29]:

Table 3
Design parameters of solar evacuated tube collector [30].
Parameters

Values

Absorber plate area per tube AP
Insolation area per tube As
Diameter of glass tube Dg
Out diameter of evaporation section of heat pipe De
Out diameter of condensation section of heat pipe Dc
Length of condensation section Lc
Coating absorptivity a

Coating emissivity ep
The transmissivity of glass tube s
The emissivity of glass tube eg
_
Working fluid mass flow rate m
The number of tubes g

0.175 m2
0.19 m2
100 mm
8 mm
14 mm
60 mm
0.92
0.08
0.9
0.88
0.03 kg/s
8

X

X

mi À

mi wi À

Qþ

X

mo ¼ 0

X

mi hi À

mo wo ¼ 0
X

mo ho ¼ 0

ð18Þ
ð19Þ
ð20Þ

The simulation was developed using EES commercial version
6.883-3D, which utilizes simple programming rules and comprehensive thermophysical property functions for classical thermodynamic calculations, is especially applicable for the modeling of
cyclic process and helps researchers. Meanwhile, thermophysical
properties of the LiBr-H2O solution can be directly obtained from
the EES software.
The average relative errors between the simulated parameters
and the reference values in [29] are shown in Table 4. It can be seen
that the average relative errors of the enthalpy, concentration,
pressure and temperature in a cycle are small. The average relative
error of the mass flow rate appears to be somewhat larger, but considering its small magnitude, which varies from 0.0034 to
0.248 kg/s, the average relative error is acceptable, which validates
the veracity of the absorption chiller model. By simulating, the COP
of absorption chiller is obtained as 1.058 while slightly higher than

the reference value which is 0.9402.
To evaluate the overall energetic performance of the chiller, the
coefficient of performance (COP) can be calculated as

COP ¼
Fig. 4. A comparison of the experimental and simulated values.

X

Qc
;
Q exh þ Q hw þ W p

ð21Þ


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J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

Fig. 5. Schematic of the dual-source powered absorption chiller.

Table 4
Average relative error of each parameter.
Average relative error
m (kg/s)
12.71%

h (kJ/kg)
0.08%

W (%)
0.80%

P (kPa)
0.07%

T (°C)
0.57%

Ignoring the power consumption of the solution pump, the COP
depends strongly on the ratio of the hot water heat to the exhaust
gas heat. Based on the view of reference [29], an analysis was conducted on the influence of the heating proportion on the COP and
cooling output, as shown in Fig. 6. It can be seen that the COP is
approximately 1.20, and the cooling output reached a maximum
when the heat inputs of the hot water and exhaust gas were the
same, which presents a consistent conclusion with the reference
simulation. The scene which hot water heat to the exhaust gas heat
ratio was 1:1.0 was assumed to be the base work condition. If the
hot water or exhaust gas inputs decrease relative to the base work
condition, both cooling outputs decrease. However, the COP always
increases with a decreasing ratio of hot water and exhaust gas, and,
moreover, an increase in the exhaust gas causes the COP to
increase more quickly than seen for the hot water. This indicates
that an increase in the proportion of the exhaust gas has a greater
contribution to the COP than that of hot water.
4. Performance evaluation criteria
The energy efficiency (gee) and exergy efficiency (gee) were
employed to evaluate the thermodynamic performance of the
hybrid CCHP system. The systematic energy efficiency, gee, is

Fig. 6. Influence of the proportion of heating sources on the COP and cooling
output.

where Qc is the produced cooling (kW), Qexh and Qhw are the
exhausted heat from the exhaust gas and hot water (kW), respectively, and Wp is the power consumption of the pump (kW).

gee ¼

E þ Qc þ Qh
Q biomass þ Q solar

ð22Þ

To measure the contributions of the solar and biomass subsystems,
the subsystem primary energy ratios, ge,sol and ge,bio, are

gee;sol ¼ Q c;sol =Q solar and

ð23Þ


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J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

À

Á

gee;bio ¼ E þ Q c;bio þ Q h =Q biomass ;

ð24Þ

where E (kW) is the electrical output, Qc (kW) is the total cooling
output, and Qc,sol and Qc,bio are the two parts of the total cooling output from the solar subsystem and biomass subsystem, respectively.
Qh (kW) is the hot water output, Qbiomass (kW) is the biomass energy
input, and Qsolar (kW) is the solar energy input. Qsolar and Qbiomass
comprise the overall energy input, and the proportion of the two
parts will influence the system energy efficiency.
The system exergy efficiency, gex, is

gex ¼






À 1 Q c þ 1 À TT 0 Q h

 h
:
EX biomass þ 1 À TT 0 Q solar
T0
T rw

Eþ

ð25Þ

sol

Similarly, the subsystem exergy efficiencies, gex,sol and gex,bio, are

gex;sol ¼

0


T0
T0
Q solar and
À 1 Q c;sol
1À
T rw
T sol



gex;bio ¼ E þ

ð26Þ


 !0

T0
T0
Qh

EX biomass ;
À 1 Q c;bio þ 1 À
T rw
Th



ð27Þ

where T0 is the reference ambient temperature, EXbiomass is the
exergy of the biomass input, and T rw ; T h and T sol are the mean temperatures of the refrigerated water, domestic hot water and solar
collector, respectively. Likewise, the proportion of the solar energy
input to biomass energy input affects the exergy efficiency. The reference state in the exergy analysis was defined as 101.325 kPa and
25 °C.
Since the biomass and solar energy are pollution free, the
hybrid system which integrated these two resources in this paper
may show a significant performance in carbon emission. In order to
evaluate the systematic environmental effects, carbon emission
reduction ratio (CERR) is introduced. The reference system is a typical biomass-fired Organic Rankine Cycle-CCHP system for the
same products which produces electricity by the turbine through
organic rankine cycle utilizing the biomass combustion heat and
recovers the condensing heat for heating and cooling. In order to
obtain the primary energy consumption of ORC-CCHP system,
coefficient of performance (gT) is selected as the evaluation criteria,
when organic medium is R245fa and the ejector coefficient of
which is at its maximum, the coefficient of performance is 0.53
[32]. Due to the consumed primary energy of these two system
are the same, here only needs to compare the biomass consumption when calculates CERR. Therefore, CERR can be calculated as:
EþQ c þQ h

CERR ¼

gT gb

À Q biomass

ð28Þ

EþQ c þQ h

gT gb

where gT is the performance coefficient of ORC-CCHP part, gb is the
efficiency of biomass-fired boiler, which is 0.85 [33], and Qbiomass is
the consumed biomass energy.

5. Results and discussion
5.1. System integrated design case
A building with a 100 kW electricity load was used as the design
condition of the hybrid CCHP system and as a case study. The collected heat and outlet temperature of the evacuated collector were
set to match that of the recovered jacket water. Therefore, to make
the outlet temperature of the solar hot water close to that of the
jacket water, three collectors were connected in series. Assuming
that the inlet water temperature was 70 °C, using the collector
model in Section 3.3.2, the outlet water temperatures of the three
collectors in turn were calculated to be T 1o ¼ 76:1  C;T 2o ¼ 82:1  C,
and T 3o ¼ 87:9  C. Thus, the heat collection of the basic collector
column could be determined. Because the solar hot water heat
was assumed to match the jacket water heat, the rows of the basic
collector column could be determined. Furthermore, the total collector area under the design condition was calculated to be 96 m2

for an 800 W/m2 solar irradiance. Using the thermodynamic models in Section 3 and the design parameters in Table 1, the results at
the design work condition are summarized in Table 5.
At the design condition, when the absorption chiller was only
driven by solar hot water, the calculated energy efficiency of the
solar subsystem (ge,sol) was 47%. Similarly, when the absorption
chiller was only driven by the recovered heat from the ICE, the
energy efficiency of the biomass subsystem () was calculated to
be 61%. Therefore, as the participation of the solar subsystem
increases, the system energy efficiency decreases. The solar subsystem exergy efficiency (gex,sol) and biomass subsystem exergy
efficiency (gex,bio) were 9.4% and 6.22%, respectively. Therefore, an
increase in biomass energy or a decrease in solar energy input
improved the system exergy efficiency. However, in terms of the
system products, an increase in the solar energy input can produce
more solar hot water, which will result in a higher cooling output,
while keeping the other parameters constant. In general, high solar
energy often occurs at noon in the hot summer when the cooling
demand just reaches its maximum. The adoption of a solar subsystem can decrease the demand for power for electrical refrigeration
and relieve the load fluctuations on the ICE. Furthermore, it
reduces biomass consumption.
5.2. Performance analysis with variable work conditions
The system does not always run under design work condition,
thus it is necessary to explore the system performance under different off-design conditions for the timely response to different
external variations. In this paper, variable external conditions
include electricity load factor (5–100%) and solar irradiation (0–
900 W/m2) are discussed. When change any one of these two variations, the other inherent system conditions and external conditions keep its design state unchanged. Meanwhile, it is important
to note here that this paper merely concentrated on the system

Table 5
Results at the design work condition.
Parameter

Value

Parameter

Value

Solar irradiance
Solar collector area
Mean temperature of solar collector
Inlet temperature of exhaust gas
Solar energy input
Biomass consumption
Heat of exhaust gas
Heat of jacket water
Heat of solar hot water

800 W/m2
96 m2
127 °C
459.7 °C
76.8 kW
613.7 kW
82.15 kW
51.29 kW
45.23 kW

Cooling
Heating
Power

System energy efficiency
Solar system energy efficiency
Biomass system energy efficiency
Exergy efficiency
Solar system exergy efficiency
Biomass system exergy efficiency

197.2 kW
102.7 kW
100 kW
57.9%
47.0%
61.0%
16.1%
9.4%
16.2%


574

J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

performance analysis under variable conditions but not specific to
any operation strategy research.

5.2.1. Variable electricity load factor
When the building electrical load changes, the performances of
the main subsystems vary as shown in Fig. 7. The performance of
the ICE can be expressed by the generation efficiency (ge), exhaust
gas heat recovery efficiency (gr,exh) and jacket water heat recovery

efficiency (gr,jw). The gr,exh and gr,jw remained basically unchanged,
but the generation efficiency rose slightly with the addition of the
electric load factor. It is easy to understand that a fixed ICE often
has a high efficiency at a high electric load factor and the overall
heat loss ratio is essentially invariant under different electric load
factors, the gr,exh and gr,jw vary slightly, and ge shows an incremental trend.
Moreover, it can be seen that the cooling output increases
almost linearly with the electric load factor. This is probably attributable to the addition of recovered heat, which is related to the
increase in the electric load factor. The COP rapidly increases at
first then increases slowly. Although a fixed ICE has an essentially
constant proportion of hot water heat to exhaust gas heat, the existence of basic solar hot water heat still can induce a change in the
proportion of the overall heating source, and thereby possibly produce a change in the COP. At the preliminary stage in the growth of
the electric load factor, an increase in the exhaust gas could greatly
increase its proportion of the heating source, which is expressed as
a change in the COP. With further increases, because of the gradually increasing proportion of hot water, there is no incremental
change in the COP.
The system energy efficiency and exergy efficiency change with
the electrical load factor as shown in Fig. 8. Their increases are also
caused by the greater contribution of the biomass subsystem. At

the preliminary stage, a low electricity output implies a low biomass input; although the rate of contribution can be greater than
that of the solar subsystem, the biomass subsystem can only play
a small part, and the values of the energy efficiency and exergy efficiency are therefore low. As the electrical output increases, more
biomass can take part in the overall performance, and the two
parameters will therefore increase rapidly. However, because of
the existence of the basic solar subsystem, a further increment will
not be able to greatly influence the two parameters. Therefore, the
energy efficiency and exergy efficiency both increase gradually at
first, and then increase slowly after reaching a certain level.

5.2.2. Variable solar irradiation
At a 100 kW electrical output, the variable solar irradiance
mainly influences the collected solar heat and solar collection efficiency (gs), and the cooling output and COP of absorption chiller
also changed, as shown in Fig. 9. When the solar irradiance was less
than 100 W/m2, the solar collectors could not collect enough heat
to recover the basic heat loss, and the curve of gs therefore began
at a critical value of approximately 100 W/m2. It can be observed
that the increasing solar irradiance improved gs, which rose
quickly in the beginning, then gradually slowed at a high irradiance. This is the case because at initial stages, an increment in solar
irradiance can deeply promote phase-change heat transfer in the
heat pipe, which will result in a holistic increment of the solar collection efficiency. When the irradiance rises further, the overall
heat transfer will finally reach saturation, and the solar collection
efficiency therefore will remain ultimately unchanged.
In addition, an increase in solar irradiance will cause the
absorption chiller to produce a greater cooling output, while
decreasing the COP. The reason is that higher solar irradiance
implies a higher solar hot water output, which increases the total
heat input of the absorption chiller, which naturally enhances
the cooling output. On the contrary, an improvement in the solar
hot water leads to a higher proportion of hot water in the heating
source, which weakens the contribution of the exhaust gas and
then decreases the COP.
The variations in the system energy efficiency and exergy efficiency are shown in Fig. 10, which shows that they trend downward in a similar manner. This is probably due to the higher
contribution of the biomass subsystem to the overall system performance relative to that of the solar subsystem. Therefore, an
increase in solar energy results in a decrease in energy efficiency
and exergy efficiency. From this perspective, if the system emphasizes increases in system energy and exergy efficiency, the
assumption of solar irradiance should not be too great.

Fig. 7. Variation in performance with the electrical load factor.

Fig. 8. System energy efficiency and exergy efficiency with electrical load factor.

Fig. 9. Influence of solar irradiance on the COP, cooling output and solar collection
efficiency.


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J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

0.61

Solar energy ratio=0.0

0.59

Energy efficiency

0.57
0.55
0.53
0.51
0.49
0.47
Fig. 10. Influence of the solar irradiance on the energy efficiency and exergy
efficiency.

0.45
0

1

2

ð29Þ

Similarly, the biomass energy ratio is defined as the ratio of the
variable biomass energy input to the nominal solar energy input:

Rb ¼

Q bio;v
;
Q sol;n

5

6

7

8

(a)
0.61
0.59
0.57

Energy efficiency

Q sol;v
Rs ¼
:
Q bio;n

4

Biomass energy ratio

5.3. Complementarity performance between biomass and solar energy
In the overall system performance, the proportions of the solar
energy and biomass energy inputs play a crucial role on the energy
efficiency and exergy efficiency. Herein, the solar energy and biomass energy ratios are defined to express their proportions in the
entire CCHP system. The solar energy ratio is defined as the ratio
of the variable solar energy input to the nominal biomass energy
input at the base design condition:

3

0.55
0.53
0.51
0.49

0.1

ð30Þ

0.47
0.45

where Rs and Rb are the solar energy and biomass energy ratios,
respectively, and Qbio,n and Qsol,n are the nominal inputs of the biomass energy and solar energy at the design condition, respectively.
Their values can be found in Table 5. Qsol,v and Qbio,v are the variable
inputs of the biomass energy and solar energy under different
conditions.
Regarding the detailed results the variable conditions, the biomass energy ratio varied from 0 to 8, which corresponds to a variation in the electrical output from 1 to 100 kW. The solar energy
ratio varied from 0 to 0.6, and the solar collector area therefore varied from 0 to 460.8 m2, which was within the maximum area of the
roof of the case study building.
5.3.1. First law of thermodynamics analysis
The variation in energy efficiency under different conditions is
shown in Fig. 11. From Fig. 11(a), it can be seen that under all of
the solar energy ratio (Rs) conditions, as the biomass energy ratio
(Rb) increased, the energy efficiency increased and then leveled.
Moreover, the energy efficiency leveled earlier with decreasing
Rs. This can be explained because a higher biomass ratio can lead
to a higher electrical output and greater proximity to the design
condition, which will result in a higher ge and an almost linear
increase in both the cooling and heating outputs. Therefore, by
analyzing the composition of formula (16), it can be concluded that
when the solar subsystem remains unchanged, an increase in the
biomass energy ratio should improve the energy efficiency. In
addition, low Rs leads to a decreased solar subsystem effect, and,
from formula (16), the energy efficiency will be greatly influenced
by Rb. Therefore, when Rs is 0, the energy efficiency is maximized
and becomes constant earlier.
Meanwhile, at certain Rb, an increase in the solar energy ratio
results in a reduced energy efficiency, and the details in the change

0

0.1

0.2

0.3

0.4

0.5

0.6

Solar energy ratio

(b)
Fig. 11. Influence of the biomass energy and solar energy ratios on the energy
efficiency.

are shown in Fig. 11(b). The figure shows that the energy efficiency
decreases steeply at a low solar energy ratio, then gradually changed slowly as the solar energy ratio further increased. This can be
explained because in the total system performance, variations in
the solar energy input can only influence the cooling output, and
although the COP of the absorption chiller was greater than 1 in
most cases, the energy efficiency still could not increase because
of the basic biomass energy input. When the Rb was high (e.g., 8),
the energy efficiency was reasonably maintained at a relatively
high level (no less than 55%). However, when the biomass energy
ratio was reduced to 0.1, the energy efficiency declined rapidly to
approximately 47% at the low solar energy ratio, and further

increases in the solar energy ratio will not improve this situation.
It can be concluded that a higher biomass energy ratio always
corresponds to higher energy efficiency, and to slow the rate in
the decline of the energy efficiency caused by an increasing solar
energy ratio, the biomass energy ratio should always be high in
system operation.
5.3.2. Second law of thermodynamics analysis
The variations in exergy efficiency under the different conditions
are shown in Fig. 12. The exergy efficiency varies from 9% to 17%,
high biomass ratio corresponds to high exergy efficiency and low
solar energy ratio conducive to the velocity of exergy efficiency


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J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

80%

Carbon emission reduction ratio

18%

Exergy efficiency

Rs=0.0
15%

12%

800W
700W
600W
500W
400W
300W
200W

60%

40%

20%

0%
0%
9%

0

1

2

3

4

5

6

7

20%

40%

60%

80%

100%

Electric load ratio

8

Biomass energy ratio

Fig. 13. Carbon emission reduction ratios under different conditions.

(a)
100%

Carbon emission reduction ratio

Exergy efficiency

18%

15%

12%

9%

80%

60%

40%

Rs=0.0
20%

0%
0

0.1

0.2

0.3

0.4

0.5

0

0.6

1

2

3

4

5

6

Solar energy ratio

Biomass energy ratio

(b)

(a)

8

100%

Carbon emission reduction ratio

Fig. 12. Influence of the biomass energy and solar energy ratios on the exergy

efficiency.

increase. It can be clearly seen that the curvilinear trend under different biomass and solar energy ratios are similar to its primary
energy ratio variation tendency under the given condition that confirms the complementary performance from two different angles.
It is parallel to aforementioned explanations that because of the
small contribution of the solar subsystem, an increase in the solar
energy ratio weakens the influence of biomass subsystem and then
results in a reduction of the exergy efficiency. In summary, higher
biomass energy ratio or lower solar energy ratio can increase the
exergy efficiency.

7

0.1
1

80%

2
60%

3
4

40%

20%

0%

5.4. Environmental analysis

0

0.1

0.2

0.3

0.4

0.5

0.6

Solar energy ratio
Compared with the biomass-fired Organic Rankine Cycle-CCHP
system, the variation in carbon emission reduction ratio under different electric ratios and solar irradiances are shown in Fig. 13.
It can be observed that when the solar collection area keeps
constant, with the increment of solar irradiance, the hybrid system
will show higher CERR. Meanwhile, when electric load ratio
increases, the values of CERR will diminished gradually and tends
to be almost uniform at full-load status. At design condition, the
CERR is about 30.87%, which shows enormous advantage in reduction ability of system which integrated biomass and solar energy.
From the perspective of energy ratio analysis, remain the solar
irradiance unchanged, the variation in carbon emission reduction

(b)
Fig. 14. Influence of the biomass energy and solar energy ratios on the CERR.

ratio under different biomass and solar energy ratios are shown
in Fig. 14.
The Fig. 14 reveals that when enlarges the solar collection area,
the increment of solar energy ratio can be benefit for the systematic carbon emission reduction. The reason can be easily understand that certain biomass energy input corresponds to settled
products of electricity, cooling and heating, when the solar energy


J. Wang, Y. Yang / Energy Conversion and Management 124 (2016) 566–577

takes part in, the cooling will raise and then enhances the coal consumption of reference separated system and will not increase the
carbon emission for the hybrid system at the same time. Otherwise, in Fig. 14(a), it can be observed that when the solar collection
area decreases to zero, i.e. Rs is 0.0, it is special that with the addition of biomass energy ratio, the CERR shows a slight upward
trend. This condition merely operates under the driven of biomass,
the upward trend shows high biomass energy ratio is benefit for
pure biomass CCHP system. And when the system integrates with
solar energy, the performance will converse and could improve the
integral level of carbon reduction to some extent.
In general, when the internal engine combustion operates from
part-load status to full-load, the solar irradiance will play less
influence on CERR. And at a certain electric load ratio, high solar
energy ratio is more conducive to emission reduction than low
ones.
6. Conclusions
This paper investigated a hybrid CCHP system driven by
biomass and solar energy and configured and integrated the
submodules of a solar evacuated collector, a dual-source powered
mixed-effect LiBr-H2O chiller and an ICE. The system thermodynamical performances under variable electricity outputs and solar
irradiances were analyzed, and the complementarity performance
between the solar and biomass energy were discussed. The following conclusions were reached.

Firstly, the biomass subsystem had a greater contribution to the
total system energy efficiency and exergy efficiency than that of
the solar subsystem. Under the design work condition, the system
energy efficiency and exergy efficiency were 57.9% and 16.1%,
respectively, the energy efficiencies of the solar and biomass subsystems were 47.0% and 61.0%, respectively, and the exergy efficiencies were 9.4% and 16.2%, respectively.
Secondly, the analysis of the variable work conditions indicated
that an increase in the electrical output corresponded to a linear
increase in the cooling output and a gradual increase in the COP
that improved the energy efficiency and exergy efficiency. However, a high solar irradiance resulted in higher cooling output,
although it decreased the COP because of the higher proportion
of low-grade heat in the dual-sources chiller, which in total
decreased the energy efficiency and exergy efficiency.
Thirdly, a high biomass energy ratio always corresponds to a
higher energy efficiency, and a decrease in the solar energy ratio
can cause an increase in the energy efficiency, which reaches a constant value earlier. Regarding the exergy analysis, an increase in
the solar energy ratio increased the exergy efficiency at higher biomass energy ratios. However, the opposite result was seen at a low
biomass energy ratio. In summary, the biomass subsystem had a
greater impact on the exergy efficiency in the hybrid CCHP system.
Last but not least, it is worth noting that the more solar subsystem take important part in hybrid system, the higher carbon emission reduction the hybrid system will be.
Acknowledgements
This research has been supported by the National Natural
Science Foundation of China (Grant No. 51406054).
References
[1] Han J, Ouyang L, Xu Y, Zeng R, Kang S, Zhang G. Current status of distributed
energy system in China. Renew Sustain Energy Rev 2016;55:288–97.
[2] Elmer T, Worall M, Wu S, Riffat SB. Fuel cell technology for domestic built
environment applications: state of-the-art review. Renew Sustain Energy Rev
2015;42:913–31.

577

[3] Han W, Chen Q, Sun L, Ma S, Zhao T, Zheng D, et al. Experimental studies on a
combined refrigeration/power generation system activated by low-grade heat.
Energy 2014;74:59–66.
[4] Roman KK, Alvey JB. Selection of prime mover for combined cooling, heating,
and power systems based on energy savings, life cycle analysis and
environmental consideration. Energy Build 2016;110:170–81.
[5] Li M, Jiang XZ, Zheng D, Zeng G, Shi L. Thermodynamic boundaries of energy
saving in conventional CCHP (Combined Cooling, Heating and Power) systems.
Energy 2016;94:243–9.
[6] Wang JJ, Jing YY, Zhang C. Optimization of capacity and operation for CCHP
system by genetic algorithm. Appl Energy 2010;87:1325–35.
[7] Ameri M, Besharati Z. Optimal design and operation of district heating and
cooling networks with CCHP systems in a residential complex. Energy Build
2016;110:135–48.
[8] Dong L, Liu H, Riffat S. Development of small-scale and micro-scale biomassfuelled CHP systems – a literature review. Appl Therm Eng 2009;29:2119–26.
[9] Mahesh A, Sandhu KS. Hybrid wind/photovoltaic energy system
developments: critical review and findings. Renew Sustain Energy Rev
2015;52:1135–47.
[10] Pérez-Navarro A, Alfonso D, Ariza HE, Cárcel J, Correcher A, Escrivá-Escrivá G,
et al. Experimental verification of hybrid renewable systems as feasible energy
sources. Renewable Energy 2016;86:384–91.
[11] Ayub M, Mitsos A, Ghasemi H. Thermo-economic analysis of a hybrid solarbinary geothermal power plant. Energy 2015;87:326–35.
[12] Jafarian M, Arjomandi M, Nathan GJ. A hybrid solar and chemical looping
combustion system for solar thermal energy storage. Appl Energy
2013;103:671–8.
[13] Wang J, Mao T, Sui J, Jin H. Modeling and performance analysis of CCHP
(combined cooling, heating and power) system based on co-firing of natural
gas and biomass gasification gas. Energy 2015;93(Part 1):801–15.
[14] Prasartkaew B, Kumar S. Experimental study on the performance of a solarbiomass hybrid air-conditioning system. Renewable Energy 2013;57:86–93.

[15] Karellas S, Braimakis K. Energy–exergy analysis and economic investigation of
a cogeneration and trigeneration ORC–VCC hybrid system utilizing biomass
fuel and solar power. Energ Convers Manage 2016;107:103–13.
[16] Hashim H, Ho WS, Lim JS, Macchietto S. Integrated biomass and solar town:
Incorporation of load shifting and energy storage. Energy 2014;75:31–9.
[17] Wang Z, Tian Y, Zhao L, Tian Y, Meng H, Hou S. Optimal design of hybrid solarassisted biomass energy heating system Transactions of the Chinese. Soc Agric
Eng 2012;19.
[18] Zhou P, Liu Q, Peng S, Jin H. A novel hybrid solar and biomass energy
complementary power generation system. J Eng Thermophys 2014.
[19] Puig-Arnavat M, Bruno JC, Coronas A. Review and analysis of biomass
gasification models. Renew Sustain Energy Rev 2010;14:2841–51.
[20] Baruah D, Baruah DC. Modeling of biomass gasification: a review. Renew
Sustain Energy Rev 2014;39:806–15.
[21] Puig-Arnavat M, Bruno JC, Coronas A. Modified thermodynamic equilibrium
model for biomass gasification: a study of the influence of operating
conditions. Energy Fules 2012;26:1385–94.
[22] Puig-Arnavat M, Bruno JC, Coronas A. Modeling of trigeneration configurations
based on biomass gasification and comparison of performance. Appl Energy
2014;114:845–56.
[23] Ptasinski KJ, Prins MJ, Pierik A. Exergetic evaluation of biomass gasification.
Energy 2007;32:568–74.
[24] Kalina J. Integrated biomass gasification combined cycle distributed
generation plant with reciprocating gas engine and ORC. Appl Therm Eng
2011;31:2829–40.
[25] Yun KT, Cho H, Luck R, Mago PJ. Modeling of reciprocating internal combustion
engines for power generation and heat recovery. Appl Energy
2013;102:327–35.
[26] Skorek-Osikowska A, Bartela Ł, Kotowicz J, Sobolewski A, Iluk T, Remiorz L. The
influence of the size of the CHP (combined heat and power) system integrated
with a biomass fueled gas generator and piston engine on the thermodynamic

and economic effectiveness of electricity and heat generation. Energy
2014;67:328–40.
[27] Skorek-Osikowska A, Bartela L, Kotowicz J, Sobolewski A, Iluk T, Remiorz L. The
influence of the size of the CHP (combined heat and power) system integrated
with a biomass fueled gas generator and piston engine on the thermodynamic
and economic effectiveness of electricity and heat generation. Energy
2014;67:328–40.
[28] Kalogirou SA. Solar Energy Engineering. 2 ed; 2014.
[29] Wang JL, Wu JY. Investigation of a mixed effect absorption chiller powered by
jacket water and exhaust gas waste heat of internal combustion engine. Int J
Refrig 2015;50:193–206.
[30] Zi-nian H, Fu-lin J, Hong-chuan G, Wei L. Study on thermal performances of
heat pipe evacuated tubular collectors. ACTA Energiae Solaris Sinica 1994;15.
[31] Shiraishi M, Kikuchi K, Yamanishi T. Investigation of heat transfer
characteristics of a two-phase closed thermosyphon. Heat Recovery Systems
1981.
[32] Zishen L, Weiyi L, Borui X, Xiangdong J, Jian J. Performance research of
combined cooling heat and power-organic rankine cycle system installed with
heat pump using mixtures. Proc CSEE 2015;35.
[33] Qiang L, Yuanyuan D, Hongwei S. Thermal performance analysis of a biomassfired organic rankine cycle. Proc CSEE 2013;33.