The process of cooling 200 T/hr of cyclohexane from 38˚C to 15˚C was simulated using Aspen Hysys to determine the volume of water and the temperature of water going to chiller as shown in Figure 1. The sensible heat of the cooling process was estimated as follows [4] :

Q s = m s C p s ( t s o − t s 1 ) (1)

where, Q_s is the rate of heat loss by the hot cyclohexane, m_s is the mass flow rate of cyclohexane, Cp_s is the specific heat capacity of cyclohexane at mean temperature, t_so is the temperature of cyclohexane after cooling and t_si is the inlet temperature of cyclohexane.

In other to estimate the water requirement, energy balance on the plate heat exchanger was done:

Q s = m w C p w ( t w o − t w i ) (2)

where, t_wo is the temperature of the water returned to the chiller evaporator after cyclohexane cooling, which is to be determined (as shown in Figure 1) and t_wi is temperature of the chilled water with a value not greater than 15˚C.

Simulation of the vapour compression cycle and cyclohexane cooling were done at evaporator temperature of 5˚C, 8˚C, 10˚C, 12˚C, & 13˚C to respectively to determine the refrigerant that will require least energy input in the compressor. The refrigerants used were 1,1,1,2-tetraflouroethane (R134a), 1,1,1,2,2- pentaflouroethane (R125), propylene (R1270) & 1,1-diflouroethane (R152a). The volume of the refrigerant required, compressor power requirement and temperature of water returned to evaporator were noted.

Using the cyclohexane cooling simulation results, the cooling load of the chiller was calculated using Equation (3) [5] .

Q e v a p = V ρ C p w ( T i n − T o u t ) (3)

The cooling load was converted to chiller capacity in refrigeration tons (TR) using a conversion factor of 1 TR is 3.5 kW [6] .

The ideal P-h diagram for the chiller is shown in Figure 2. It was used to determine the thermodynamic properties of the refrigerant at stage 1 (compressor suction), stage 2 (compressor discharge), stage 3 (condenser outlet) and stage 4 (expansion valve outlet/ evaporator inlet.

1) Compressor Design

Since the compression process is isentropic, the discharge temperature was calculated using the temperature variation between stages in isentropic process.

T d = T S [ r ( γ − 1 γ ) ] (4)

In this work, the discharge temperature does not exceed 107˚C because the discharge temperature should be less than the critical temperature of refrigerant [7] .

Using Figure 2, the compressor energy input could be determined as follows

W c = h 2 − h 1 (5)

From the definition of enthalpy, Equation (5) could be transformed as [8]

W c = m ˙ C Pr ( T 2 − T 1 ) (6)

To determine the isentropic efficiency of the designed compressor, isentropic head was estimated as [9]

W i s e n = C p R ( T 2 r e v − T 1 ) (7)

To get T^rev, the compression was assumed reversible and the entropy of isentropic process was zero [4]

Δ S = C p ln ( T 2 r e v T 1 ) − R ln ( P 2 / P 1 ) = 0 (8)

The isentropic efficiency of the compressor was calculated using Equation (9) while the compressor power was estimated using Equation (10) [10] .

η i s e n = W i s e n W s (9)

Power = m ˙ C Pr ( T 2 − T 1 ) (10)

In this study, an isentropic efficiency range of 70% - 85% was used [9] .

2) Evaporator and Condenser Design

The chiller was designed using a condenser temperature of 40˚C and evaporating temperature of 10˚C. The type of evaporator designed was flooded water-cooled shell and tube heat exchanger [11] . The tube side fluid was water while the refrigerant, R134a, was the shell side fluid. The condenser was shell and tube.

Taking energy balance on the evaporator, the heat transfer rate can be estimated as shown on Figure 2.

Q e v a p = m ( h 1 − h 4 ) (11)

The area of evaporator required for heat transfer was estimated using Equation (12) while the number of tubes was estimated using Equation (14) [12] .

Q e v a p = U A Δ T (12)

The range of values of “U’’ in Equation (12) in this study was 1080 - 3600 kJ/h m²・C (300 - 1000 W/m²・K) for evaporator [13] and 3600 - 14,400 kJ/h m²・C (1000 - 4000 W/m²・K) for condenser [11] .

The log mean temperature difference (LMTD) was evaluated as shown below [11] .

Δ T = T D 1 − T D 2 ln ( T D 1 T D 2 ) (13)

N t = A π d o 2 (14)

Using Figure 2, the heat transferred in the condenser can be established by taking energy balance at the condenser [14] .

Q c o n d = m ( h 2 − h 3 ) (15)

3) Metering Device

At the valve, the refrigerant exists as saturated liquid (f) and saturated vapour (g) due to the flashing action. The dryness fraction of the refrigerant was denoted by “X”. To get the dryness fraction of the refrigerant at the valve outlet, the enthalpy of the refrigerant was obtained as [7]

h 4 = h f + x 4 ( h g − h f ) (16)

where, h_f and h_g were gotten from chart (16)

From Figure 2, the coefficient of performance of the chiller can be determined as:

C O P = h 1 − h 4 h 2 − h 1 (17)

For this study, the value of COP was considered to be greater than 5.7 since the Carnot COP of the chiller was 5.7.

Figure 4 shows how the compressor power input vary with evaporator temperature for all the refrigerants used. It indicates that R134a was the best fit refrigerant for the design since it had the least compressor power input at each respective evaporator temperature.

Increased in evaporator temperature from 5˚C through 13˚C reduced the compressor power input for all the refrigerants due to drop in compression ratio which reduced the compressor work as shown in Figure 4.

This effect was maximum when R134a was used; thus, R134a is best fit for the design.

Evaporating temperature of less than 5˚C was not considered to avoid freezing of the chilled water inside the tubes of evaporator; while evaporating temperature greater than 13˚C was not considered to avoid temperature cross.

Figure 5 shows the effect of compressor duty on coefficient of performance for the different refrigerants used. The best fit refrigerant for the design was 1,1,1,2-tetrafluoraethane (R134a) since it required the least compressor power as shown on Figure 5. A chiller with less power input had high coefficient of performance and is cost effective [6] . R134a required a compressor power of 378.1 KW at 10˚C while R152a required 390 kW compressor power at 10˚C since the discharge temperature was higher when R152a was used.

Figure 5 also show that the coefficient of performance of a chiller increased with decreased in compressor power because coefficient of performance is the ratio of quantity of energy absorbed in evaporator to the energy input of the refrigeration cycle. R134a has the highest coefficient of performance of 6.49 since it had the least compressor input of 333.6 kW as shown in Figure 5.

Table 3 shows the compressor sizing results gotten from Hysys. The power input of the designed compressor was 378.6 kw while it had adiabatic efficiency of 75%. The compressor had an adiabatic head of 1995 m. This high power consumption was as a result of increase in evaporator duty and refrigerant flow. The maximum discharge temperature of the designed compressor was satisfactory as it was less than the critical temperature of refrigerant which is 107˚C. The compressor was single stage because the compressor ratio was less than 4.0 [11] .

Table 4 shows that a design heat duty of 7,799,004 kJ/hr was required in the evaporator of this vapour compression cycle. The evaporator design temperature was 10˚C. This was because for a flooded water-cooled evaporator, the temperature approach, which is the difference between the evaporating temperature and the chilled water outlet temperature, should be 3.3˚C - 5.6˚C [6] . The tube side outlet temperature was 15˚C because the design was made at a worse case scenario of a temperature approach of the plate cooler, where the cooling of cyclohexane was done, at 0˚C.

The area of heat transfer was 288 m² as compared to that of defective chiller which is 244 m² (from vendor diagram) and the evaporator had 764 tubes. This area increase was as a result of the rise in heat duty.

Table 5 suggests that the condenser duty required for the design was 9,167,040 kJ/hr and a condensing temperature of 40˚C was used because as a rule of thumb, the condenser temperature is 10˚C - 15˚C above ambient temperature [15] .

The overall heat transfer coefficient of the condenser was 3796 kJ/h-m² C and the area of heat transfer was 332.5 m² as compared to that of the defective chiller that was 319 m². The total number of tubes of the designed condenser was 882.

To avert the effect of corrosion, carbon steel was used for shell design.

Table 6 shows the expansion valve sizing results. For this design, thermostatic expansion valve was considered. The valve has a flow coefficient (C_v) of 46.17 gpm at 50% opening.

The flow rate through the valve was 5.22 × 10⁴ kg/h while the pressure drop of the valve was 597.1 kPa.