Analysis of a Vapor Absorption Machine to Air

7/28/2019 Analysis of a Vapor Absorption Machine to Air

http://slidepdf.com/reader/full/analysis-of-a-vapor-absorption-machine-to-air 1/23

Project Undertaken for

Automobile Engineering (MEE 428)

2012-2013

Analysis of a Vapor AbsorptionMachine to Air-Condition the Cabin of a

Car, using the exhaust gas heat from

the Diesel Engine.

By: Amna Nashit (10BEM0064)

Shadab Khan (10BEM0103)

Kanav (07BEM)

Project guide: Prof. Ramesh Kumar C



Abstract:

On an average, 30-32% of the total power produced by a diesel engine is lost as heat of the exhaust gases. A vehicle

consumes 15% extra fuel when air conditioning is switched ‘ON’ [2]. Thus, the car air conditioning system causes

approximately 0.5 l/100 km excess fuel consumption on average [1]. A study of vapor absorption system, which runs

the a/c using the waste heat generated, is imperative to a better fuel economy. It not only eliminates the need of an

engine driven compressor but also reduces the temperature of the exiting flue gas. Also it increases the gas mileage

by 5 to 10% as a consequence (an SAE study).

An analytical study was conducted on a vapor absorption machine, which uses LiBr-water pair as an absorber-

refrigerant system, to find the change in COP of the system with load. The absorption system is assumed to be

attached to a 10hp, 1500rpm four stroke compression ignition engine. The exhaust gas was used to heat water to

temperatures of 48.5 oC at a pressure of 2.9343 kPa. The super heated steam was then used to act as a refrigerant fo

our absorption machine. This system is acceptable since the generator temperature needed for a LiBr-water pair in a

single effect vapor absorption machine is 75-120oC (Fundamentals, ASHRAE-1993). The COP was found to increase

proportionally with load.

Introduction:

In the vapor absorption system, a physicochemical process replaces the mechanical process of the vapor compression

system by using energy in the form of heat rather than mechanical work. The main advantages of this system lie in

the possibility of using the energy in hot waste gases. This study investigates the performance of a vapor absorption

unit utilizing the waste heat in the exhaust gases from a diesel engine that is used to represent the main propulsion

unit of a road vehicle. If such a system could be used to provide air conditioning, there would be no need for the IC

engine and the compressor of the vapor compression system. There would also be a corresponding reduction in

exhaust gas pollution.

Besides optimum fuel economy, another concern which has emerged in the last five to ten years is the search for

environmentally-benign refrigerants and techniques. Wide-spread efforts are currently underway to develop

replacements for the traditionally used halogenated hydrocarbon refrigerants, which contribute to ozone depletion

and greenhouse warming. Here yet again, the vapor absorption model provides an elegant solution.



Selection of absorbent-refrigerant pair:

The materials that meet the refrigerant absorbent pair must meet the following criteria [ASHRAE, 1993]:

1. Absence of solid phase: The refrigerant-absorbent pair must not form a solid phase over the range of

composition and temperature to which it might be subjected.

2. Volatility ratio. The refrigerant must be much more volatile than the absorbent so the two can beeasily separated.

3. Affinity. The absorbent should have strong affinity for refrigerant under working conditions. This

affinity (1) causes a negative deviation from Rault’s Law and results in an activity coefficient of less

than unity for the refrigerant; (2) reduces the amount of absorbent to be circulated and consequentlythe waste of thermal energy from sensible heat; (3) reduces the size of liquid heat exchanger that

transfers heat from absorbent to pressurized refrigerant-absorbent solution in practical cycle.

4. Pressure. Operating pressures, largely established by the physical properties of the refrigerant,should be moderate.

5. Stability. Almost absolute chemical stability is required, because fluids are subjected to severe

conditions over many years of service. Instability can cause undesirable formation of gases, solids, or

corrosive substances. 6. Corrosion. Since the fluids or substances created by instability can corrode materials used in

constructing equipments, corrosion inhibitors should be used.

7. Safety. Fluids must be non-toxic and non-inflammable if they are in an occupied dwelling.

8. Transport properties. Viscosity, surface tension, thermal diffusivity and mass diffusivity are all

important characteristics of the refrigerant-absorbent pair. 9. Latent heat. Refrigerant’s latent heat should be high so that circulation rate of solution can be kept at

minimum.

No known refrigerant-absorbent pair meets all requirements listed. But LiBr-water pair has been found to contain

several advantages over others. It’s high volatility ratio, high affinity, high stability, high safety and high latent heat

make it ideal for our system. Also, less pump work is needed compared to other units due to operation at vacuum

pressures. Evaporator temperatures are limited to above 0 degrees as the refrigerant (water) freezes above that

which is ideal for our cause (air-conditioning).

Although, the tendency of crystallization of LiBr salt at moderate concentrations ( >0.65 kg LiBr/Kg of solution) and

the systems need to be designed in hermetically sealed units since they operate at vacuum pressures do pose a

challenge.



Absorption System:

Fig 1. Single effect LiBr-water absorption cycle



Cycle description and assumptions:

State 1: High pressure, high temperature pure refrigerant (here water vapor) enters the condenser.

State 2: Water vapor loses heat to the surrounding, in an isobaric process, in the condenser and turns into low

temperature, high pressure liquid.

State 3: After passing through the throttle valve, the fluid undergoes isenthalpic expansion. The final state is low

temperature, low pressure liquid.

State 4: Heat gain in the evaporator is an isobaric process. It results in the change of state of the refrigerant from low

pressure liquid to low pressure gas.

State 5: High concentration LiBr solution (low concentration in water) is sprayed over the incoming water vapor inthe absorber. The resulting solution is a low concentration LiBr solution (high concentration in water). Heat is

produced during this exothermic absorption process which is continuously extracted by cooling water circulation.

State 6: High pressure weak solution (high concentration of water)

State 7: Temperature increases from state 6

State 8: Very high temperature and highly pressurized strong solution of LiBr (weak in water)

State 9: Temperature is lower than that of state 8. Constant pressure heat exchanger.

State10: High temperature, low pressure strong LiBr solution (weak in water).



Psychometric calculations:

For the Capacity of the system:

1. Optimum temperature for an air-conditioned space is 20-27 o C.

2. RH = 40-60%

3. Air speed = 15m/min

We have performed the following design calculations:

P=1atm, T1=32o C RH=80%

V1=10m3/min (Data based on average conditions for Oct, 2012 over Vellore)

T2=20o C RH=50%

Plotting the data on a p-h chart:



H1=97kJ/kg of dry air H2=31kJ/kg of dry air

w1 = 0.225 kg water/kg of dry air w2 = 0.08 kg water/kg of dry air

hw (at T=9o

C) = 37.8kJ/kg v1= 0.91m3/kg

vel =15m3/min

ma = vel/v1 = 16.5kg/min

mw =ma (w1-w2) = 2.9kg/min

Q out= ma (h1-h2)-mw*hw =16.323kW

This is approx. 4.6Tons of refrigeration

The Q out found is used as Q e in the above calculations

Equations used:

A flow analysis of the system is carried out with the following assumptions:

i. Steady state and steady flow

ii. Changes in potential and kinetic energies across each component are negligible

iii. No pressure drops due to friction

iv. Only pure refrigerant boils in the generator.

The circulation ratio (λ) is defined as the ratio of strong solution flow rate to refrigerant flow rate. It is given

by:

λ = mss/m

This implies that the strong solution flow rate is given by:

mss = λ m

The analysis is carried out by applying mass and energy balance across each component.

Condenser:

m1 = m2 = m

Q c = m (h1 − h2)



Pc = Psat (Tc)

where Tc is the condenser temperature

Expansion valve (refrigerant):

m2 = m3 = m

h2 = h3

Evaporator:

m3 = m4 = m

Q e = m (h4 − h3)

Pe = Psat (Te)

where Te is the evaporator temperature

Absorber:

From total mass balance:

m + mss = mws

mss = λ m ⇒mws = (1 + λ) m

From mass balance for pure water:

m + (1 − Xss ) mss = (1 −Xws ) mws

⇒ λ =Xws/Xss − Xws

Q a = m h4 + λ m h10 − (1 + λ) m h5

or, Q a = m [(h4 − h5 ) + λ(h10 − h5 )]

Solution pump:

m5 = m6 = mws

WP = mws (h6 − h5) = (1 + λ) m (h6 − h5)

However, if we assume the solution to be incompressible, then:

WP = (1 + λ) m vsol (P6 − P5) = (1 + λ) m vsol (Pc − Pe)

where vsol is the specific volume of the solution which can be taken to be approximately equal to 0.00055 m3/kg.



Solution heat exchanger:

m6 = m7 = mws

m8 = m9 = mss

Heat transfer rate in the solution heat exchanger, Q HX is given by

Q HX = (1 + λ) m (h7 − h6) = λ m (h8 − h9)

Generator:

m7 = m8 + m1

Heat input to the generator is given by:

Q g = m h1 + λ m h8 − (1 + λ) m h7

or,

Q g = m [(h1 − h7) + λ (h8 − h7)]

Solution expansion valve:

m9 = m10 = mws

h9 = h10

The COP of the system is given by:

COP =Q e / (Q g + WP) ≈ Q e / Q g

Solution heat exchanger:

εHX=(T7-T6)/(T8-T6)



To determine analytically the complete state of the system we need:

1. Operating temperatures

2. Weak and strong solution concentrations

3. Effectiveness of pre-heater

4. Refrigeration capacity

If any 3 out of the above mentioned properties is known, the remaining properties can be found from the given

graphs or the stated equations.

Refrigerant temperature

A0 = 2.00755 A1=0.16976 A2= -3.133362e-3 A3= 1.97668e-5;

B0= 124.937 B1= -7.71649 B2=0.152286 B3=-7.95090e-4;

0C

Pressure

C = 7.05; D = -1596.49; E = -104095.5;

P = exp( C + D / (T’+273) + E / (T’+273)2

) kPa

Enthalpy

F0 = -2024.33 F1= 163.309 F2= -4.88161 F3= 6.302948e-2 F4= -2.913705e-4 G0 = 18.2829 G1= -1.1691757 G2= 3.248041e-2 G3= -4.034184e-4 G4= 1.8520569e-6 H0 = -3.7008214e-2 H1= 2.8877666e-3 H2= -8.1313015e-5 H3= 9.9116628e-7 H4= -4.4441207e-9

kJ/kg



Fig. 3 Duhring plot for LiBr-water solution





Design parameters for the single effect LiBr –water absorption cooler

Parameter Symbol Example-value

Capacity Q e 16.323 kW

Evaporator temperature T4 6 0C

Generator solution exit temperature T8 90 0C

Weak solution mass fraction XWS 55% LiBr

Strong solution mass fraction XSS 60% LiBr

Solution heat exchanger exit temperature T9 65 0C

Generator (desorber) vapor exit temperature T1 85 0C

Data for single effect LiBr –water cooling systemPoint # h (kJ/kg) _mm (kg/s) p (kPa) T (

0C) X (%LiBr) Remarks

5 83 0.053 0.934 34.9 55

6 83 0.053 9.66 34.9 55

7 145.4 0.053 9.66 65 55 Sub-cooled liquid

8 212.2 0.0486 9.66 90 60

9 144.2 0.0486 9.66 54.8 60

10 144.2 0.0486 0.934 44.5 60

1 2628 0.0044 9.66 85 0 Superheatedsteam

2 185.3 0.0044 9.66 44.3 0 Saturated liquid

water 3 185.3 0.0044 0.934 6 0

4 2511.8 0.0043 0.934 6 0 Saturated vapor



Tests on the diesel engine:

Performance tests were carried out on the diesel engine and the results tabulated as shown below.

The Greek alphabet

SI.

No.

Weight

kg

Speed

rpm

Time

sec

TFC

g/min

Torque

Nm

B.P.

kW

SFC

Kg/kWh

FP

kW

IP

kW m%

Heat

I/P

kW

BT

%

1 0 1552 31.65 15.5 0 0 ∞ 4.25 4.25 0 11.47 0

2 0.9 1520 26.31 18.7 8.89 1.398 0.8 4.25 5.648 24.75 13.838 10.1

3 1.3 1506 20.91 2.35 18.879 2.964 0.475 4.25 7.214 41 17.39 17.04

4 2.1 1500 18.69 2.63 27.993 4.395 0.359 4.25 8.645 50.84 19.462 22.58

Table 1

SI.

No.

Load

kg

Speed

rpm

Manometer

Reading

(mm)

Time for

10cc of fuel

t, sec

Time for 1L

of cooling

water, sec

ma

g/s

mf

g/s

Exhaust

temp 0C

Exhaust loss

kJ/hr

1 0 1552 255 31.65 19.44 14.4 0.26 214 9754.7

2 0.9 1520 245 26.31 19.44 14.1 0.31 252 11575.39

3 1.3 1506 240 20.91 19.44 14.0 0.39 304 14235.53

4 2.1 1500 235 18.69 19.44 13.8 0.44 346 16291.85

Table 2

Design of Generator:

To design the generator and study the dependence of the mass flow rate of exhaust gas with the COP of the

absorption machine, the following assumptions have been made.

Assumptions:

1. The composition of exhaust gas and air is considered to be nearly the same and values of transportproperties of dry air at 1atm are used.

2. It is assumed, as can be argued, that pool boiling dominates in the generator and pool boiling

equations for zeotropic liquids are used to determine heat transfer coefficient.

A MATLAB code was written to plot mass flow rate of exhaust gas vs COP of absorption system and the same has

been listed in appendix 1. The theory and equations behind the model can be described as follows.



Transport properties of air [7]:

All transport properties are calculated at the bulk temperature which can be roughly stated as the mean of the entry

and exit temperature of the exhaust gases.

Tb = (t1+t2)/2 K

Density

kg/m

3

Specific heat

J/kg K

Dynamic viscosity

N.s/m2

Thermal conductivity

W/m K

Prandtl number

Reynold’s number

Based on whether the flow is laminar or turbulent (i.e. ,Re is less than or greater than 2500, respectively), Nusselt

number is found from which the heat transfer coefficient can be found using the following relation.

Heat transfer coefficient

W/m

2K

Transport properties of LiBr solution :

There exist only limited studies on the thermodynamic and transport properties of LiBr solution despite of it being

such a popular refrigerant. Below is a compilation of such studies that can completely define the state of LiBr

solution, at any given temperature and concentration.

Thermodynamic properties have already been listed above ( eq )



Thermal conductivity [8]

W/m K

a1 = -1407.5255 a2 = 11.051253 a3 = -1.467414e-2 b1 = 38.98555 b2 = -0.24047484 b3 = 3.4807273e-4 c1 = -0.26502516 c2 = -1.5191536e-3 c3 = -2.3226242e-6 Density [8]

kg/m

3

Dynamic viscosity [8]

A1 = (0.494122 + 1.63967X - 1.4511X

2

) x 10

3

A2 = (2.86064 - 9.34568X + 8.52755X2) x 10

4 A3 = (0.703848 - 2.35014X + 2.07809X

2) x 10

2

Now, in a generator the normal equations that apply to heat exchangers break down as the Reynold’s number and

Nusselt number equations are valid, by definition, only to flows where either sensible heating or latent heating is

involved. As in a generator both sensible and latent heat exchange takes place, it was the author’s opinion that pool

boiling equations would be more suitable for such a scenario. Having said that, the author would also like to point out

that this is an a-typical case of boiling as the concentration of solution in vapor phase and liquid phase are different

and so are their temperatures. Here we deal not only with heat but mass transfer effects. Mass transfer tends to

reduce nucleate boiling heat transfer coefficients and, in some cases, may reduce the value of the heat transfer

coefficient by up to 90%. Detailed reviews of mixture boiling are given by Thome and Shock (1984) and by Collier and

Thome (1994).The mass transfer effect on bubble growth can be explained in simple terms as follows. Since the equilibrium

composition of the more volatile component is larger in the vapor phase than in the liquid phase, the more volatile

component preferentially evaporates at the bubble interface, which in turn reduces its composition there and

induces the formation of a diffusion layer in the liquid surrounding the bubble. The partial depletion of the more

volatile component at the interface increases that of the less volatile component, which increases the bubble point

temperature at the interface. This incremental rise in the local bubble point temperature can be denoted as∆Ө.

Hence, to evaporate at the same rate as in a pure fluid, a larger superheat is required for a mixture.

The effect of mass transfer on nucleate pool boiling heat transfer can therefore be explained by introducing the

parameter ∆Ө, which represents the increase in the bubble point temperature at the surface due to preferential

evaporation of the more volatile component. At a given heat flux, the boiling superheat of the mixture is ∆TI+∆Ө while that for an ideal fluid with the same physical properties as the mixture is ∆TI. Thus, the ratio of the mixtureboiling heat transfer coefficient αnb to that of the ideal heat transfer coefficient αnb,I at the same heat flux is [9]:

The value of ∆TI is the wall superheat that corresponds to αnb,I, which is determined for instance using the Cooper

correlation with the molecular weight and critical pressure of the mixture. Hence, as the value of ∆Ө increases, the

ratio αnb/αnb,Idecreases, which means that a larger wall superheat is required in a mixture to transfer the same heat



flux. As exploited in an early mixture boiling prediction by Thome (1983), the maximum value of ∆Ө is the boiling

range of the mixture ∆Ө bp, which is equal to the difference between the dew point and the bubble point

temperatures at the composition of the liquid.

Starting from a mass transfer balance around an evaporating bubble and simplifying with an approximate slope of the

bubble point curve, the following expression was obtained to predict heat transfer in the boiling of mixtures

where βmL is the mass transfer coefficient in the liquid (set to a fixed value of 0.0003 m/s). The value of αnb,I is

determined with one of the pure fluid correlation suggested by Cooper (1984):

Overall heat transfer coefficient U can be given by

And then the overall heat transferred in the generator is given by

where ∆Tlm is the logarithmic mean temperature in 0C and A in the inner surface area of the pipe in m2.

Result and conclusion:

The following plots were obtained and conclusions drawn.It can be clearly concluded that the COP of the absorption machine, and hence the efficiency of the air-conditioning

system attached to the vehicle, greatly improves as the vehicle gains speed. This is owing to the fact that the

temperature of exhaust gases increase with the increase in speed. Also, the mass flow rate decreases as the speed

increases as can be seen from plot 2. This increases the net time of heat exchange and hence increases the COP. But

this relationship is not linear as can be concluded from the graph and verified analytically from the stated equations.

It is also seen that a well designed heat exchanger can provide enough heat to the LiBr solution to keep the system

running even in no load conditions, i.e., stalling. It can be thus concluded that a vapor absorption machine is an

exploitable mode of running air conditioning in the vehicle and further work must be undertaken to design and

model such a system.



Plot 1 : Variation of COP with Load (kg)

Plot 2: Variation of mass flow rate of exhaust gas (kg/s) with load (kg)



Plot 3: Variation of COP with the mass flow rate of exhaust gas (kg/s)

Appendix 1 – MATLAB code used

fid=fopen('data.txt'); M = fscanf(fid,'%g %g %g %g %g %g %g %g %g %g %g', [11,inf]); fclose(fid);

M = M';

for k=1:length(M) t1 = M(1,i); %Exhaust gas entry temperature

t2 = M(2,i); %Exhaust gas exit temperature t3 = M(3,i); %Temperature of LiBr weak solution entering the generator t4 = M(4,i); %Temperature of LiBr strong solution exiting the generator Mex = M(5,i); %Mass flow rate of exhaust gas m3 = M(6,i); %mass flow rate of weak solution into the generator m4 = M(7,i); %mass flow rate of Strong solution from the generator m5 = M(8,i); %mass flow rate of refrigerant from the generator Di = M(9,i); %inner diameter of pipe Do = M(10,i); %outer diameter of pipe L = M(11,i); %Length of pipe



%Exhaust gas Heat transfer coefficient Tb = (t1+t2)/2; %Bulk temperature of exhaust gas %Properties of exhaust gas

%Density rhoex = (351.99/Tb) + (344.84/(Tb^2));

%Specific heat Cpex = 1030.5 - (0.1997*Tb) + (3.9734e-4 * (Tb^2));

%Dynamic viscosity MUex = (1.4592e-6 * (Tb^1.5)) / (109.1 + Tb);

%Thermal conductivity Kex = (2.334e-3 * (T^1.5)) / (164.54 + Tb) ;

%Prandtl number Pr = (Cpex*MUex)/Kex;

Rex = (4*Mex) / (pi*Di*MUex); %Reynold's number

%Finding Nusselt number if (Rex<2300) %Laminar flow

if (((L/Di)/(Rex*Pr))<0.01) Nu = 1.67*(((Rex*Pr)/(L/Di))^0.03);

else

Nu=3.66; end

else %Turbulent flow if((Rex>2500)&&(Rex<1.25e6)&&(Pr>0.6)&&(Pr<100))

Nu = 0.023*(Rex^0.8)*(Pr^0.4); elseif (((L/Di)>2)&&((L/Di)<20))

Nu = Nu (1+((Di/L)^0.7)); end

end

Hi = (Nu*Kex)/L; %Exhaust gas side heat transfer coefficient

% Lithium Bromide side heat transfer coefficient

Xs = 60; %Mass fraction of strong solution of LiBr Xw = 55; %Mass fraction of weak solution of LiBr T = t4; X = Xs;

%Properties of LiBr

%Thermal conductivity a1 = -1407.5255; a2 = 11.051253; a3 = -1.467414e-2; b1 = 38.98555; b2 = -0.24047484; b3 = 3.4807273e-4; c1 = -0.26502516; c2 = -1.5191536e-3; c3 = -2.3226242e-6;



At = a1+(a2*T)+(a3*(T^2)); Bt = b1+(b2*T)+(b3*(T^2)); Ct = c1+(c2*T)+(c3*(T^2));

Klb = At + Bt*X + Ct*(X^2);

%Density rho = 1.40818 - 0.713995*X + 2.64232*(X^2) - ((0.12318 + 0.946268*X)*T)/1000;

%Dynamic Viscosity A1 = (0.494122 + 1.63967*X - 1.4511*(X^2))*e3; A2 = (2.86064 - 9.34568*X + 8.52755*(X^2))*e4; A3 = (0.703848 - 2.35014*X + 2.07809*(X^2))*e2; MUlb = exp(A1 + (A2/T) + A3*log(T));

%Refrigerant temperature A = 2.00755; 0.16976; -3.133362e-3; 1.97668e-5; B = 124.937; -7.71649; 0.152286; -7.95090e-4;

for i=1:3 s1 = B(i)*(Xî); s2 = A(i)*(Xî);

end t5 = (T-s1)/s2;

%Pressure C = 7.05; D = -1596.49; E = -104095.5; P = exp( C + (D / (t5+273)) + (E / ((t5+273)^2)));

%Enthalpy F = -2024.33; 163.309; -4.88161; 6.302948e-2; -2.913705e-4;

G = 18.2829; -1.1691757; 3.248041e-2;-4.034184e-4; 1.8520569e-6; H = -3.7008214e-2; 2.8877666e-3; -8.1313015e-5; 9.9116628e-7; -4.4441207e-9; %Enthalpy of weak solution coming into the generator h3=0; for i=1:5

h3= h3 + ( F(i)* (Xwî) + (G(i) * (Xwî)) * t3 + (H(i) * (Xwî)) * (t3^2)); end %Enthalpy of strong solution leaving the generator h4=0; for i=1:5

h4= h4 + ( F(i)* (Xsî) + (G(i) * (Xsî)) * t4 + (H(i) * (Xsî)) * (t4^2)); end %Enthalpy of vapour leaving the generator tau = (t5+273)/674.14; pr = P/22.064; h5 = (10258.8 - (20231.3/tau) + (24702.8 / (tau^2)) - (16307.3 / (tau^3))...

+ (5579.31 / (tau^4)) - (777.284 / (tau^5)) + pr*((-355.878 / tau)... + (817.288 / (tau^2)) - (845.841 / (tau^3))) - (pr^2)*(160.276/(tau^3))... + (pr^3)*((-95607.5 / tau) + (443740 / (tau^2)) - (767668 / (tau^3))... + (587261 / (tau^4)) + (167657 / (tau^5))) + (pr^4)*((22542.8 / (tau^2))... - (84140.2 / (tau^3)) + (104198 / (tau^4)) - (42886.7 / (tau^5))));

%heat capacity



a = 3.462023; b = 2.679895e-2; c = 1.3499e-3; d = -6.55e-6; Cp = (a + (b*X) + (T*(c + (d*X))));

%Dew point temperature Z = [ -1.313448e-1, 1.820914e-1, -5.177356e-2, 2.827426e-3, -6.380541e-5,

4.340498e-7; 9.967944e-1, 1.778069e-3, -2.215597e-4, 5.913618e-6, -7.308556e-8, 2.788472e-6;

1.978788e-5, -1.779481e-5, 2.002427e-6, -7.667546e-8, 1.201525e-9, -6.641716e-

12]; Z = Z'; Tdp=0; for i=1:6

for j=1:3 Tdp = Tdp + (B(i,j)*(Xî)*(T^j));

end end

Prlb = (Cp*MUlb)/Klb; %Prandtl number q = (m5*h5) + (m4*h4) - (m3*h3); %Heat absorbed by the solution

%heat transfer coefficient for LiBr solution side of the heat exchanger %using pool boiling equations Ho1 = 55 * (Prlb^0.12) * ((-0.4343 * log(Prlb))^0.55) * (M^0.5) * (q^0.67);

%accounting for mass transfer due to azeotropic solution beta = 0.0003; %mass transfer coefficient Ho = Ho1 * ((1 + ((ho1 / q) * (Tdp-T) * ( 1 - exp ( -q / (rho*h4*beta)))))^-1);

%Over-all heat transfer coefficient U = ( Hi^-1 + Rfi + (Di/(2*Ksteel))*log(Do/Di) + (Di/Do)*Rfo + (Di/Do)*(Ho^-1))^-1;

%Average heat transfer in a generator A = pi*Di*L; %Inner surface area for heat exchange Tlm = ((t1-t3)-(t2-t4))/log((t1-t3)/(t2-t4)); %logarithmic mean temperature

difference Q(i) = U*A*Tlm;

end

plot (Mex,Q); %Since COP depends directly on Q if capacity is kept constant

Plots:

1. Power input vs Exhaust gas temperature

2. Input power vs COP

3. Exhaust gas temperature vs COP



References:

1. Contribution of the Air Conditioning System to Reduced Power Consumption in Cars, Thomas E. J.

Heckenberger et al- Behr GmbH & Co. KG , 2008 Paper Number: 2008-21-0047

2. Automotive Air Conditioning Systems with Absorption Refrigeration , Joseph R Akerman, 1971 Paper Number

710037

3. Fundamentals, ASHRAE Handbook, 1993

4. Vapor absorption refrigeration in road transport vehicles. I Horuz

5. A car air-conditioning system based on an absorption refrigeration cycle using energy from exhaustgas of an internal combustion engine, Wang et al

6. Design and construction of LiBr-water absorption machine. Florides et al, 2002 Energy Conversion

and Management 44 (2003) 2483 – 2508.

7. Expressions for Thermophysical properties of superheated steam. V.I. Lachkov et al, Measurementstechniques Vol.42 No.1, 1999

8. Thermophysical property data for Lithium Bromide water solutions at elevated temperatures.

ASHRAE, 1991.

9. Fundamentals of boiling on tubes and tube bundles. J.R Thome, Swiss Federal Institute of Technology, Lausanne.

10. Thermodynamic property data for Lithium bromide-water solutions at high temperature. Y. Kaita,

International Journal of Refrigeration 24(2001) 374-390 11. Mc Neely LA. Thermodynamic properties of aqueous solutions of LiBr. ASHRAE Transactions

1979. 85 (Part 1):413-34

12. Rockenfeller U. Laboratories result: solution = Libr-H2O, properties = P-T-X, heat capacity.Unpublished data. Boulder City (NV): Rocky Research Inc. 1987.

13. Iyoki S, Uemura T. Heat capacity of the water-lithium bromide system and the water-lithium

bromide-zinc bromide-lithium chloride system at high temperatures. International Journal of

Refrigeration 1989; 2379-88.

Documents

Analysis of a Vapor Absorption Machine to Air