Solar and biomass are the two main renewable energy sources extremely suitable
for drying application. Since the availability of solar is limited only during
the sunny days, the uses of biomass burner as a backup heater is highly preferable.
Many agriculture producers nowadays are applying a solar dryer assisted by electricity
or biomass fuel to extend the period of drying after sunset and during the cloudy
days, as well as drying large capacity of products. Conventional fuel operated
driers are more efficient but it is beyond the reach of rural people with limited
product (Prasad et al., 2006). High prices and
shortages of fossil fuels have increased the emphasis on using alternative renewable
energy resources (Muhlbauer, 1986).
Biomass fuel is suitable to be applied in burner as a backup heater in drying
application. Biomass can be described as stored solar energy (www.aesenergy.net).
Biomass sources include food crops, grassy and woody plants, residues from agriculture
or forestry, organic components of municipal and industrial wastes and animal
waste. The combustion products from complete combustion of biomass generally
contain nitrogen, water vapor, carbon dioxide and surplus of oxygen. If there
is a surplus of solid fuel due to incomplete combustion, the products of combustion
are combustible gases like Carbon monoxide, CO, Hydrogen, H2 and
traces of Methane and non-useful products like tar and dust (Rajvanshi,
1986). These combustible gases are not suitable for drying food application.
So the usage of the gas to gas heat exchanger (G-to-G HEX) is crucial to make
sure that the gas that will be used for food drying application is clean, free
from smoke, soot and ash in order to protect the food from being contaminated.
The biomass fuel combustion heated the wall of the G-to-G HEX and allowed the
transmission of the heat to the ambient air, resulting in temperature raise.
This process produces a heated clean air for food drying application.
A compact G-to-G HEX needs large heat transfer areas on both fluid sides which
can be done by adding fins. The fins extend the heat transfer surfaces and promote
turbulence (Wang et al., 1999). The higher the
thermal conductivity of the material used, the higher the heat transferred across
the heat exchanger. A lot of studies which focused on the design of burner with
heat exchanger have been done by previous researchers. Fang
(2000) studied on a clamshell heat exchanger in residential gas furnace.
The clamshell is created by joining two steel panels containing half channel
in a manner similar to joining the two halve of a sandwich. The half channels
are stamped or molded into the panels forming complete gas channels at assembly
to guide the flow of combustion gases inside the channel. The furnace heat exchanger
separates the high temperature flue gas stream from the low temperature circulating
air stream while at the same time transferring thermal energy from the former
to the latter. The heat exchanger is made of aluminized steel. The density increases
and its volumetric flow rate decreases as the flue gas is cooled while travel
Tomimura et al. (2004) studied on a multi layered
type of gas to gas heat exchanger using porous media. The porous metal plate
used in this process is extremely high porosity material and give a large temperature
drop and larger amount of converted radiant energy. This type of heat exchanger
had higher overall heat transfer coefficients then the conventional heat exchanger
and also exhibit excellent reaction characteristics as a steam reformer. The
enthalpy of the high temperature gas is effectively transferred to the porous
metal plate via extremely high heat transfer coefficient between the flowing
gas and the porous plate with fine mean pore size, at the same time a substantially
large surface area of the porous for heat transfer.
Al-Omari (2006) focused on a biomass furnace design
for experimental and investigation on combustion and heat transfer characteristics.
A digital weighing scale is used to monitor and record the fuel mass in the
bed during the combustion process. LPG combustion was found to be more convenient
and effective to initiate the combustion around the 2 to 3 min which is stabilized
and guided by means of conical bluff body.
The use of biomass burner as back up heater is relatively synonym in the industry
of drying. A suitable design of controllable biomass burner is important so
that the required heat is supplied to dryer without affecting the quality of
the product to be dried. Madhlopa and Ngwalo (2007) studied
on an indirect solar dryer with biomass backup heaters. The biomass burner was
made of brick and consists of rock pebbles which acted as a thermal storage.
Thanaraj et al. (2004) came out with a furnace
which consists of heat exchanger using bricks, clay and cement to the rotary
dryer. The same material type of burner also has been reported by Prasad
(2006) Prasad and Vijay (2005), Tarigan
and Tekasakul (2005) and Bena and Fuller (2002).
Ocused on a solar hybrid tunnel dryer incorporated with a biomass stove-heat
exchanger, consists of a cross-flow shell and tube heat exchanger. Serafica
and del Mundo (2005); (Bhattacharya et al., 2000)
focused on a biomass gasifier stove design as a backup heater to the hybrid
solar dryer for fish and fruits and vegetables, respectively. The biomass gasifier
consists of shell and fin heat exchanger configuration and the heat delivery
and combustion rate could be controlled using a butterfly valve at the primary
air inlet. Among the biomass fuel materials that has been reported in biomass
burner application are coconut shells (Serafica and del
Mundo 2005), woodchips (Bhattacharya et al.,
2000; Madhlopa and Ngwalo, 2007), charcoal (Prasad
and Vijay, 2005), paddy husk (Thanaraj et al.,
2004), fuel wood (Prasad et al., 2006; Bena
and Fuller, 2002; Tarigan and Takasakul, 2005) and
briquetted rice husk.
The El Paso Solar Energy Association in 2010 provides a basic guidelines to
dry food where the temperature ranges between 37 to 71°C will effectively
kill bacteria and inactivate enzyme although temperatures around 43°C are
recommended for solar dryers and aims to remove 80 to 90% of moisture from the
food. The allowable maximum temperature of heat under solar or biomass burner
supplied for most of the tropical fruits, vegetables and also fish (Serafica
and del Mundo, 2005) drying is about 70°C. For safe storage, crops usually
dried to a final moisture content of < 14% with equilibrium moisture content
= 14% and RH of 80-90% is preferred (Ayensu, 1997).
The present study is documenting the design procedure of burner/gas-to-gas
heat exchanger as a backup heater for solar drying.
DESIGN OF THE DRYING SYSTEM
The flow diagram of biomass burner with gas-to-gas heat exchanger is shown
in Fig. 1. The unit is uniquely designed to meet the drying
To dry food, a heat exchanger is needed so that the flue gas can be separated
from the clean warm air to protect the food from contamination by the smoke,
soot and ash. Waste drying product does not require any specific temperature
limit and quality control, thus the heat from direct fuel can be used as the
source of heat. The material to be dried is located inside the drying chamber
of solar dryer.
||Biomass burner with gas to gas heat exchanger
|| Conceptual design condition
The maximum allowable temperature in the drying chamber either under solar
or heat from burner is 65°C. This temperature is selected based on the studies
of solar drying and biomass backup heater design journals and furthermore, it
is the suitable drying temperature for all types of product.
The average measured temperature of the solar dryer system under solar mode
for 8 h day-1 is shown in Table 1.
Based on Table 1, the recorded drying temperature inside
the drying chamber was 49°C. In order to increase the rate of drying and
dryer efficiency, the use of biomass burner as a backup heater is favorable
to maintain temperature inside the drying chamber within the range of 50 to
MATERIALS AND METHODS
Conceptual design: Basically, the unit consists of three zones. The
lower zone is for the solid biomass burning. The upper part is the gas-to-gas
heat exchanger. The upper part has a cylindrical shape compromises two zones;
inner cylindrical hot zone and outer annulus cold zone. The flue gas flow up
in the inner zone and the cold air flow up in the outer zone, similar to parallel
double pipe HEX. The thermal process in the unit is that the burned biomass
produces hot flue. Previous measurements have shown that the flue temperature,
Tf is around 320°C. The flue moves up within the inner cylindrical
passage by natural draft assisted by a chimney. During its up flow, the hot
flue exchanges heat with air. Eight holes with 25.4 mm drilled on the outer
surface to permit ambient air to flow to the outer zone of the G-to-G HEX. The
air enters the annulus outer zone of the heat exchanger at ambient temperature
(~30°C) and moves up due to buoyancy effect assisted by a chimney. To enhance
the heat transfer process, 8 extended surfaces have been added in both sides
of the cylinder. The heat required, Q obtained from the dryer design is used
in order to design the heat exchanger inside the burner.
Detailed calculation procedure: Among the steps involved are finding
of thermal flue coefficient, hflue to be applied in the thermal balance
equation, finding of thermal air coefficient, hair, deliberating
the amount of mass flow rate from the dryer into design calculation of the burner
and finally, comparison of the calculated length characteristic, Lc
by iteration where the parameter setting of error is 0.001. The overall design
of the burner is shown in Fig. 2.
||The preliminary unit design and the parameter assumption
Equations 1 and 2 are used to calculate
the heat supplied from the heat exchanger and the outlet temperature of flue
is the mass flow rate obtained from dryer (kg s-1), cpa is
the specific heat capacity of air (J kg-1.K), Tao is the
outlet air temperature (K) and Ta is the ambient temperature (K).
In the flue flow side,
where, pf is the density of flue (kg m-3), A is the area
of heat exchanger (m2), Vf is the velocity of flue (m
s-1), cp is the specific heat capacity of flue (J kg-1.K),
Tfi is the inlet flue temperature (K) and Tfo is the outlet
flue temperature (K).
The type of flow for flue is determined by calculating Re number as in Eq.
|| Laminar flow
|| Turbulent flow
The Nu equation is basically depends on the mode of convection heat transfer,
whether it is natural or forced. This is based on ratio of Gr/Re2.
The nature of convection heat transfer can be determined:
||Natural convection+Forcedconvection, i.e. combined
GrL is determined between the difference of mean flue temperature,
Tf and inside wall temperature, Twi of heat exchanger
using Eq. 4.
The combine Nu is evaluated by Eq. 5.
For natural convection (6) is used, as:
For laminar, C = 0.59, n = ¼; and for turbulent, C = 0.10, and n = 1/3.
Rayleigh number, RaL is
where, Pr is the Prandtl number.
For forced convection and laminar flow, the Nu is evaluated by two assumed
cases. First assuming fully developed flow of the flue; then Nu = 3.66. For
the case of considering the flow at entrance conditions, the suitable correlation
for Nu number for entry length is obtained from Eq. 8 and
9. For turbulent, fully developed forced convection, the well
known Dittus-Boelter equation has been used:
The convection heat transfer coefficient for flue, hflue is then
calculated using (10):
|| Heat transfer network from the flue to the air
The illustration of heat transfer across the wall of heat exchanger is shown
in Fig. 3.
The characteristic length, Lc or height of the heat exchanger is
obtained by deliberating the same amount of the calculated heat required from
The heat transfer from the flue to the inside wall:
Across the wall of heat exchanger:
And from the outside wall of the cylinder to air:
where, Af is the flue side thermal area (m2), A is the
air side thermal area (m2), hair is the air convection
coefficient (Wm-2.K), hflue is the flue convection coefficient
(Wm-2.K), ro is the outer radius of cylinder (m), ri
is the inner radius of cylinder (m), Tair is the mean air temperature
(K), Tf is the mean flue temperature (K) and Two is the
wall out temperature (K).
The steps of procedure are repeated again using the previous calculations across
the wall of heat exchanger. The type of flow for air is determined by calculating
Re number as shown in Eq. 15 using properties at air temperature.
where, V is the mean air velocity (m s-1) and Dh is the
hydraulic diameter (m).
GrL number for air is calculated using (17) between Two
and Tair. The calculation is followed by finding of the nature of
convection heat transfer based on Gr/Re2.
where, hair is obtained by substituting the values Nu which depending
on the type of flow into Eq. 10 by using properties of air
annulus hydraulic diameter. The Rayleigh number, RaL is obtained
from Eq. 7 which involves GrL from Eq.
4 and Pr of air.
By adapting the same criteria of 0.1< Gr/Re2 <10, the convection
heat transfer mode is checked and the relevant Nu equation is used to calculate
hair. For forced convection heat transfer, Dattus-Boelter relation
is used as shown in Eq. 18. For the case of natural convection,
the relation in Eq. 6 is used (Incropera,
where, n = 0.3 for cooling and n = 0.4 for heating. If the case is combined,
thus the Nu is evaluated as in Eq. 5.
The total area, A of heat exchanger is calculated using Eq.19.
Since the temperature distribution along the length of outer fins, Lf
is not constant, thus it is divided by 2.
where: Acylinder = Lc*π*D2 , and Afins
Finally, the new length characteristic, Lc is calculated by
using the same Q calculated from dryer shown in Eq. 20 followed
by iteration until it reached the parameter setting of error, 0.001. The purpose
of iteration is to find the corrected value based on the value of initial assumption.
Material selection of the biomass burner: The burner consists of cylinder
which is the main part of the gas heat exchanger.
|| Gas to gas heat exchanger matrix decision table
|Criteria : A: Melting point; B: Corrosive; easiness; D: Thermal
conductivity; E: Price choice : 1: Very bad; 2: Medium; 3: Good; 4: Very
good; 5: Perfect
It has a function as a major heat transfer surface, and also as a barrier to
prevent flue gas to mix with the clean ambient air. Since the burner will experience
a high combustion temperature, the selected material must have a good characteristic
in the required criteria, which are melting point, corrosive, assembly easiness,
thermal conductivity and price.
For the design, five types of material are consider reliable for the process.
Table 2 represents the decision matrix leading to the final
choice. All the materials are rating between 0 and 5 to each cell of the table,
bound to the influence of the corresponding.
According to Table 2, material that has been selected for
the burner is stainless alloy, AISI 304 with thermal conductivity, k of 14.9
W m-1. K. Even though the thermal conductivity is low, it the most
suitable type of material which can extend high combustion temperature, easy
to fabricate and excellent for joining (welding).
RESULTS AND DISCUSSION
The final design of the unit: The final boundary conditions of the unit
design are shown in Table 3. The mathematical modeling for
the governing equations for the unit designed has been converted into programming
language which is MATLAB. The programming is divided into two parts, the flue
side is to calculate hflue and the air side across the heat exchanger
is to calculate the correct height, Lc. The execution of the
program involved many iteration processes.
The initial characteristic length, Lc is assumed to be 0.4 m and
the finalize length is obtained from the iteration until the difference of error
between the new and old length reaching approximately to 0. The result of iteration
is shown in Fig. 4. The calculated Lc that obtained
at final iteration was 0.33 m.
The calculated heat required, Q from dryer design was 154 Watt and this value
is used by deliberating with burner design calculation.
|| Boundary conditions of the unit
|| Characteristic length, Lc versus number of iteration
||The designed length of the heat exchanger inside the burner
|| The components of the biomass burner
The convective flue heat transfer coefficient, hflue calculated
from the first part was 3.41 W m-2.K. In the second part, the calculated
Re was 319.17 which indicate that the flow is laminar with air convective heat
transfer coefficient, hair of 4.78 W m-2.K.
The result of design calculation of the burner at final iteration is shown
in Table 4. The value of hflue obtained from calculation
was compared with the typical values of the convection heat transfer coefficient
for gases (Incropera, 2007).
||The measured air temperature obtained from different burning
rate of wood chip
||The measured air temperature obtained from different burning
rate of rice husk
||The measured air temperature obtained from different burning
rate of EFB
For free convection of gases, the range is from 2 to 25 W m-2.K
which means that the calculated value of hflue, 3.41 W m-2.K
is valid. The final Lc is to be multiply by safety factor of 1.2
to compensate for the losses to the exchanger surroundings.
|| Result of burner design calculation
Thus, the final Lc calculated for the design of burner is 0.4 m.
The illustration of the final length and the components of the unit are shown
in Fig. 5 and 6, respectively.
Validation test of the design procedure is carried out by experimental measurements.
Various types of solid fuel have been used to compare and select the most suitable
type for the drying requirements. Wood chips, rice husk and EFB of the palm
oil have been fed at various rate, kg h-1. The measurement results
are shown in Fig. 7, 8 and 9
for wood chip, rice husk and EFB, respectively.
The analyses of measurement results are demonstrating that the designed and
fabricated unit is capable to produce the required floe rate and temperature
of the hot to conduct the drying. Rice husk shows lower performance, but still
within the required range of drying air temperature. The tested feeding rates
of the three solid biomass fuels produced higher than the air outlet design
temperature, which is 80°C. this means, that smaller feeding rate is sufficient
to provide the required drying hot air.
A thermal unit is designed and fabricated to backup a solar dryer. The unit
comprises of two main parts, biomass burner and gas-to-gas heat exchanger. The
conceptual design and the boundary conditions are based on the drying requirements
of 2.5 kg of EFB. The results from the adopted material selection criteria show
that stainless alloy, AISI 304 is the most suitable selection for the burner
and the heat exchanger unit. The mathematical calculations results show that
heat exchanger of 0.4 m height and 0.4 m diameter is sufficient to produce the
required flow rate and temperature of drying air.
This study is funded by Universiti Teknologi PETRONAS under the graduate assistantship
scheme. The advices and supports by technicians of the Mechanical Engineering
Department are highly appreciated.