Transcript IBO-Presentationversion

btrc.masdar.ac.ae
Optimal Control of Chiller Condenser Sub-cooling, Compressor
Speed, Tower Fan and Pump Speeds, and IGV
Omer Qureshi, Hassan Javed & Peter Armstrong, June 2013
1
Presentation Outline
Introduction
SCADA and Heat Balance Analysis
Component Models
Chiller System Solver
Optimization
Conclusion and Future Work
2
Introduction
Plant under consideration-(4x2500T).
Collection and analysis SCADA
Development of sub models for Individual chiller components
Validation of model
Development of solver- to execute these sub models and predict
chiller performance.
Optimize the model to produce set of conditions for optimum
power consumption.
3
District Cooling Plant
Selected District cool Plant
Capacity (4x2500T)
Shell and tube Evaporator and Condenser
Constant speed single stage centrifugal compressor
Capacity control by Pre-rotation vanes
Surge control Variable geometry diffuser
2-cell cooling tower each with variable speed fan (Fan diameter: 8m)
Variable speed chilled water pump
Constant speed condenser water pump
4
Chiller Unit
1. Maintenance manual of York Chiller(Source: Tabreed)
5
SCADA & Heat Balance Analysis
6
Components Models—Chiller Unit
Steady-state models based on first principle
Inputs
Component engineering parameters
SCADA Data
Simple models, less computation time
Four Component models for district cooling plant
Evaporator Model----Shell and tube
Condenser Model----Shell and tube
Centrifugal Compressor Model (Isentropic work + loss Mechanism)
• Constant speed
• Variable speed
Variable speed pump model
7
Evaporator Model
ENGINEERING PARAMETERS
Tubes
Length of shell
Copper
6.6 m
Tube Pass (water)
2
Total no. of tubes
1234
Tube Diameter
0.75" or 1.905x10-2 m
Tube thickness
0.028" or 7.11x10-4 m
Assumptions:
No pressure drop considered for refrigerant side
Thermal resistance from refrigerant side is neglected.
8
Evaporator Model
𝑸𝒕,𝒆
𝑯𝒍𝒊𝒒,𝒂𝒔
𝑽𝒆
𝑻𝒘,𝒊𝒏,𝒄
𝒅𝑻𝒆𝒔𝒉
𝒎𝒓
Evaporator
𝑇𝑤,𝑜𝑢𝑡,1
Evaporation
𝑇𝑤,𝑜𝑢𝑡,2𝑎
Evaporation
𝑇𝑤,𝑜𝑢𝑡,2𝑏
𝑻𝒆
𝑻𝒘,𝒐𝒖𝒕,𝒆
Superheating
𝑷𝑬𝒏𝒈𝒈.
1st Pass
2nd Pass
Two regions for refrigerant were modeled:
Evaporation
Superheating
𝞮 – NTU Method
Single Stream HX for evaporation
Crossflow HX for super heating
9
Evaporator Model
Equations utilized in Evaporator Model
𝑇𝑤,𝑜𝑢𝑡,𝑒 = 𝑇𝑤,𝑖𝑛,𝑒 −
ℎ𝑖𝑛,𝑒 = 0.023 𝑅𝑒𝑒
0.8
𝑃𝑟0.4
𝑘𝑤
𝑄𝑡,𝑒
𝑐𝑝,𝑤 𝑚𝑤
𝑈𝐴𝑒 =
𝐷𝑒,𝑖
𝐶𝑚𝑖𝑛
=
1
1
𝐴𝑖𝑛,𝑒 ℎ𝑖𝑛,𝑒 + 𝑅𝑃,𝑒
min[𝑐𝑝,𝑤 𝑚𝑤,𝑒 , 𝐶𝑝,𝑟 𝑚𝑟 ]
𝐴𝑖𝑛,𝑒,1 = 𝜋 𝐷𝑒,𝑖 𝐿𝑒 (𝑁𝑒 /2) 𝐴𝑖𝑛,𝑒,1𝑎 = 𝜋 𝐷𝑒,𝑖 𝑥𝑒 𝐿𝑒 (𝑁𝑒 /2) 𝐴𝑖𝑛,𝑒,1𝑏 = 𝜋 𝐷𝑒,𝑖 (1 − 𝑥𝑒 ) 𝐿𝑒 (𝑁𝑒 /2)
Evaporation
Evaporation
Superheating
10
Evaporator Model
Equations utilized in Evaporator Model
Equation for regressed length:
𝐿𝑒 = 8.947𝑥10−3 𝑚𝑟2 − 3.6279𝑥10−1 𝑚𝑟 + 7.227
Equation for temperatures:
𝑇𝑤,𝑜𝑢𝑡,1 = 𝑇𝑤,𝑖𝑛,𝑒 − (𝑇𝑤,𝑖𝑛,𝑒 − 𝑇𝑒 ) 1 − 𝑒 −𝑁𝑇𝑈𝑒1
𝑇𝑤,𝑜𝑢𝑡,2𝑎 = 𝑇𝑤,𝑜𝑢𝑡,1 − (𝑇𝑤,𝑜𝑢𝑡,1 − 𝑇𝑒 ) 1 − 𝑒 −𝑁𝑇𝑈𝑒,2𝑎
𝑇𝑤,𝑜𝑢𝑡,2𝑏 = 𝑇𝑤,𝑜𝑢𝑡,2𝑎 − 𝜀2𝑏 (𝑇𝑤,𝑜𝑢𝑡,2𝑎 − 𝑇𝑒 )
𝜀2𝑏 = 1 − 𝑒𝑥𝑝
1
𝐶𝑟
𝑁𝑇𝑈𝑒2𝑏
0.22
exp −𝐶𝑟 𝑁𝑇𝑈𝑒,2𝑏
0.78
−1
11
Evaporator Model
1. Maintenance manual of York Chiller(Source: Tabreed)
12
Evaporator Model
4
0.2096 C
NRMS
0.0319
3.5
Modeled Te (C)
RMS
Measured Te (C) vs Modeled Te (C)
Measured Te (C)
15% error line
-15% error line
3
2.5
2
1.5
1.5
2
2.5
3
Measured Te (C)
3.5
4
13
Condenser Model
ENGINEERING PARAMETERS
Tubes
Length of shell
Tube Pass (water)
Total no. of tubes
Sub-cooling Section:
Tube Diameter
No. of tubes
Tube thickness
Tube Surface Area
Condensation & de-superheating Section:
Tube Diameter
No. of tubes
Tube thickness
Tube Surface Area
Copper
6.6 m
2
937
0.75" or 1.905x10-2 m
180
0.028" or 7.11x10-4 m
66.78 m2
1" or 2.54x10-2 m
757
0.035" or 8.89x10-4 m
376.44 m2
Assumptions:
No pressure drop considered for refrigerant side
Thermal resistance from refrigerant side is neglected.
14
Condenser Model
𝒎𝒓
𝑻𝒄
𝑽𝒄
Condenser
𝑻𝒘,𝒊𝒏,𝒄
𝑻𝒄𝒐𝒎𝒑,𝒐𝒖𝒕
𝑷𝑬𝒏𝒈𝒈.
𝑻𝒘,𝒐𝒖𝒕,𝒄
𝑇𝑤,𝑚𝑖𝑥,𝑐
Condensation
𝑇𝑤,𝑜𝑢𝑡,2𝑎
Condensation
𝑇𝑤,𝑜𝑢𝑡,2𝑏
Desuperheating
𝑸𝒄
𝑯𝒍𝒊𝒒
Sub-cooling
1st Pass
2nd Pass
Three regions for refrigerant were modeled:
Sub-cooling
Condensation
De-Superheating
𝞮 – NTU Method
15
Condenser Model
Equations utilized in Condenser Model
1a. Sub-Cooling Section(First Pass):
ℎ𝑖𝑛,𝑐,1𝑎 = 0.023 𝑅𝑒𝑐1 0.8 𝑃𝑟 0.4
𝑈𝐴𝑐,1𝑎 =
1
1
𝐴𝑖𝑛,𝑐,1𝑎 ℎ𝑖𝑛,𝑐,1𝑎
+ 𝑅𝑃,𝑐,1𝑎
𝑇𝑐𝑠 = 𝑇𝑐2 −
𝑇𝑤,𝑜𝑢𝑡,1𝑎
𝜀𝑐,1𝑎 =
𝑘𝑤
𝐷𝑐1,𝑖
1 − 𝑒 −𝑁𝑇𝑈𝑐,1𝑎
1−𝐶𝑟,1𝑎
1 − 𝐶𝑟,1𝑎 𝑒 −𝑁𝑇𝑈𝑐,1𝑎 (1−𝐶𝑟,1𝑎)
𝜀𝑐,1𝑎 𝐶𝑚𝑖𝑛,1𝑎 (𝑇𝐶2 − 𝑇𝑤,𝑖𝑛,𝑐 )
𝑚𝑟 𝑐𝑝,𝑟
𝑚𝑟 𝑐𝑝,𝑟 (𝑇𝐶2 − 𝑇𝑆𝐶 )
= 𝑇𝑤,𝑖𝑛,𝑐 +
𝑚𝑤,𝑐 𝑥1,𝑎 𝑐𝑝,𝑤
16
Condenser Model
1b. Condensation Section (First Pass):
ℎ𝑖𝑛,𝑐,1𝑏 = 0.023 𝑅𝑒𝑐1𝑏 0.8 𝑃𝑟 0.4
𝑈𝐴𝑐,1𝑏 =
1
1
𝐴𝑖𝑛,𝑐,1𝑏 ℎ𝑖𝑛,𝑐,1𝑏 + 𝑅𝑃,𝑐,1𝑏
𝑇𝑤,𝑜𝑢𝑡,1𝑏
𝑘𝑤
𝐷𝑐2,𝑖
𝑁𝑇𝑈𝑐,1𝑏 =
𝑈𝐴𝑐,1𝑏
𝐶𝑚𝑖𝑛,𝑤
𝑥𝑐𝑎 𝑚𝑟 (𝐻𝐶2 − 𝐻𝐶3 )
= 𝑇𝑤,𝑖𝑛,𝑐 +
𝑚𝑤,𝑐 (1 − 𝑥1𝑎 )𝑐𝑝,𝑤
Mixing Section:
𝑇𝑤,𝑚𝑖𝑥,𝑐 =
𝑚𝑤,𝑐 𝑥1,𝑎 𝑇𝑤,𝑜𝑢𝑡1,𝑎 − 𝑚𝑤,𝑐 (1 − 𝑥1𝑎 ) 𝑇𝑤,𝑜𝑢𝑡1𝑏
𝑚𝑤,𝑐
17
Condenser Model
2a. Condensation Section (Second Pass):
ℎ𝑖𝑛,𝑐,2𝑎 = 0.023 𝑅𝑒𝑐2𝑎 0.8 𝑃𝑟 0.4
𝑈𝐴𝑐,2𝑎 =
1
1
𝐴𝑖𝑛,𝑐,2𝑎 ℎ𝑖𝑛,𝑐,2𝑎 + 𝑅𝑃,𝑐,2𝑎
𝑇𝑤,𝑜𝑢𝑡,2𝑎 = 𝑇𝑤,𝑚𝑖𝑥,𝑐 +
𝑘𝑤
𝐷𝑐2,𝑖
𝑁𝑇𝑈𝑐,2𝑎 =
𝑈𝐴𝑐,2𝑎
𝐶𝑚𝑖𝑛,𝑤
𝑥𝑐𝑏 𝑚𝑟 𝑐𝑝,𝑟 (𝐻𝐶2 − 𝐻𝐶3 )
𝑚𝑤,𝑐 𝑐𝑝,𝑤
2b. De-superheating Section (Second Pass):
𝑇𝑤,𝑜𝑢𝑡,2𝑏 = 𝑇𝑤,𝑜𝑢𝑡2𝑎 +
𝑚𝑟 𝑐𝑝,𝑟,2𝑏 (𝑇𝐶1 − 𝑇𝐶2 )
𝑚𝑤,𝑐 𝑐𝑝,𝑤
18
Condenser Model
0.0949 C
NRMS
0.0225
Measured Tc (C) vs Modeled Tc (C)
Measured Tc (C)
2.5% error line
-2.5% error line
32
30
Modeled Tc (C)
RMS
34
28
26
24
22
22
24
26
28
30
Measured Tc (C)
32
34
19
Condenser Model
35
0.6481 C
NRMS
0.1471
30
Modeled Tw,out (C)
RMS
Measured Tw,out (C) vs Modeled Tw,out (C)
Measured Tw,out (C)
5% error line
-5% error line
25
20
20
25
30
35
Measured Tw.out (C)
20
Compressor Model
Integral and mathematically most complex part of chiller
Constant and variable speed compressor model
Non-Dimensional loss model based on Aungier(2000)
Assumptions
Centrifugal Compressor Specification
•
Gear efficiency is taken as 90%
Refrigerant
•
Velocity profile is assumed as constant,
along the hub and tip
Rating (Btuh)
2500
•
The kinetic energy of refrigerant
entering the diffuser is completely
converted to useful energy
Rating (kW input)
1817
Rating discharge pressure (psig)
162
•
Diffuser and IGV losses are not
modeled
Rating suction pressure psig)
34
•
Water flow rate for motor cooling is
taken as constant
•
Complex engineering parameters in
impeller geometry
Rating suction temperature (F)
R134A
33/34
Impeller diameter (outlet diameter) m
0.7
Impeller hub diameter (inlet diameter)
0.3
Impeller Blade Angle (degree)
45/50
21
Compressor Model-Inputs and Outputs
Constant Speed Model
Input
IGV Positions
Constant RPM
Inlet and outlet pressure of compressor
Inlet and outlet blade and velocity angles of
impeller
Impeller Inlet and outlet engineering
parameters and dimensions
Gear efficiency
Output
Compressor Power
Pressure at impeller exit
Temperature at compressor outlet
Pressure drop due to Impeller losses
Variable speed Model
Input
Mass flow rate of refrigerant
Inlet and outlet pressure of compressor
Inlet and outlet blade and velocity angles of
impeller
Impeller Inlet and outlet engineering
parameters and dimensions
Gear efficiency
Output
Compressor Power
Compressor RPM
Pressure at impeller exit
Temperature at compressor outlet
Pressure drop due to Impeller losses
22
Validation Constant Speed Compressor Model
1600
Actual Power(kW)
Model Power(kW)
Loss Power(kW)
Model Comp Power(kW)
1400
Copmressor Power (KW)
1200
1000
800
600
400
200
0
0
200
400
600
800
No. of Observations
1000
1200
1400
23
Validation Constant Speed Compressor Model
1600
108.34 KW
NRMS
0.1553
1400
1200
Model Power(kW)
RMS
Measured Power(kW) vs Model Power(kW)
Measured Power(kW)
10% Error line
-10% Error line
1000
800
600
400
400
600
800
1000
1200
Measured Power(kW)
1400
1600
24
Variable Speed Compressor Model
𝐼𝑠𝑒𝑛𝑡𝑟𝑜𝑝𝑖𝑐 𝑊𝑜𝑟𝑘 = 𝑤𝑖𝑠𝑒𝑛

𝑃1
=
 − 1 𝜌1
−1
𝑃3
𝑃1

−1
𝜂𝑔𝑒𝑎𝑟
𝑀𝑎𝑠𝑠 𝐹𝑙𝑜𝑤 𝑟𝑎𝑡𝑒 𝑜𝑓 𝑟𝑒𝑓𝑟𝑖𝑔𝑒𝑟𝑎𝑛𝑡 = 𝑚 = 𝜙2 𝐴2 𝑈2 𝜌2
RPM is calculated in an iterative process by satisfying the following equation
𝑟𝑒𝑠𝑢𝑙𝑡 = 𝑚𝑠𝑐𝑎𝑑𝑎 − 𝑚𝑐𝑎𝑙
Total Work
𝑊𝑎𝑐𝑡 = 𝑊𝑐𝑜𝑚𝑝 + 𝑊𝑙𝑜𝑠𝑠
Loss Model Calculations
𝑊𝑙𝑜𝑠𝑠 = ∆𝑃𝑡𝑟 𝑉𝑟
Total Relative Pressure Drop (Due to Losses)
𝑖
∆𝑃𝑡𝑟 = 𝑓𝑐 (𝑃𝑡𝑟 1 − 𝑃𝑠1 )
𝑖
25
Variable Speed Compressor Model-Benefits/comparison
Compressor Power (KW)
Variable Speed Compressor (KW)
Measured Compressor Power (KW)
Power (KW)
IGV Position
1504.702
44.2
Operation Conditions:
1. mr (kg/s)
2. Pout/Pin
No. of Observations
26
Impeller Loss Model
𝑉𝑚 1
𝐼𝑛𝑐𝑖𝑑𝑒𝑛𝑐𝑒 𝐿𝑜𝑠𝑠 = 1 −
𝑊1 sin(𝑚 1 )
𝑊1𝑇ℎ
𝐷𝑖𝑓𝑓𝑢𝑠𝑖𝑜𝑛 𝑙𝑜𝑠𝑠 = 0.8 1 −
𝑊1
2
𝑡𝑏1 𝑍
+
2𝑟𝑚 1 sin(𝑚 1 )
2
2
− 𝐼𝑛𝑐𝑖𝑑𝑒𝑛𝑐𝑒 𝐿𝑜𝑠𝑠
𝑆𝑘𝑖𝑛 𝐹𝑟𝑖𝑐𝑡𝑖𝑜𝑛 𝐿𝑜𝑠𝑠 = 4𝑐𝑓
𝑊
𝑊1
2
𝐿𝐵
𝐷𝐻
(∆𝑊 𝑊1 )2
𝐵𝑙𝑎𝑑𝑒 𝐿𝑜𝑎𝑑𝑖𝑛𝑔 𝐿𝑜𝑠𝑠 =
24
( − 1)𝑉𝑚 2
𝐸𝑥𝑝𝑎𝑛𝑠𝑖𝑜𝑛 𝐿𝑜𝑠𝑠 =
𝑊1
𝐶𝑙𝑒𝑎𝑟𝑎𝑛𝑐𝑒 𝐺𝑎𝑝 𝐿𝑜𝑠𝑠 =
2
2𝑚𝐶𝐿 ∆𝑃𝐶𝐿
𝑚 1 𝑊12
2
(𝑚 𝑏 𝑊 𝑊1 )
𝐻𝑢𝑏 − 𝑆ℎ𝑟𝑜𝑢𝑑 𝐿𝑜𝑠𝑠 =
6
27
Variable Speed Compressor Model-losses profile
120
Pressure Drop (kPa)
100
80
Clearance gap loss (kPa)
Diffusion loss (kPa)
Hub-shroud Loss (kPa)
Incident loss (kPa)
Skin friction loss (kPa)
Blade Loading Loss (kPa)
Expansion Loss (kPa)
60
40
20
0
20
25
30
35
Refrigerant Mass Flow (kg/s)
40
45
50
28
Cooling Tower Model
Effectiveness NTU Method
𝐻𝑒𝑎𝑡 𝑅𝑒𝑗𝑒𝑐𝑡𝑒𝑑 = 𝑄𝑟𝑒𝑗𝑒𝑐𝑡𝑒𝑑 = 𝑚𝑤 ∗ 𝑐𝑝𝑤 ∗ (𝑇𝑐𝑤𝑠 − 𝑇𝑐𝑤𝑟)
𝐶𝑜𝑜𝑙𝑖𝑛𝑔 𝑇𝑜𝑤𝑒𝑟 𝑅𝑒𝑡𝑢𝑟𝑛 𝑇𝑒𝑚𝑝𝑒𝑟𝑎𝑡𝑢𝑟𝑒 = 𝑇𝑐𝑤𝑠 −
𝑄𝑟𝑒𝑗𝑒𝑐𝑡𝑒𝑑
𝑚𝑤 ∗ 𝑐𝑝𝑤
𝑄𝑟𝑒𝑗𝑒𝑐𝑡𝑒𝑑 =  ∗ 𝐶𝑚𝑖𝑛 ∗ (𝑇𝑐𝑤𝑠 − 𝑇𝑤𝑏 )
1 − 𝑒 −𝑁𝑇𝑈(1−)
=
1 −  𝑒 −𝑁𝑇𝑈(1−)
𝑁𝑇𝑈 =
𝑚_𝑤
𝑀𝑒𝑀
𝑚_𝑎
𝐾 ∗ 𝑎 ∗ 𝑉
𝑀𝑒𝑀 =
𝑚_𝑤
Regression Coefficient
𝑀𝑎𝑠𝑠 𝐹𝑙𝑜𝑤 𝑜𝑓 𝐴𝑖𝑟 = 𝑚_𝑎 = 𝑉𝑚𝑎𝑥 ∗ 𝜌𝑎 ∗ 𝑓
29
Cooling Tower Model
Assumptions, Specifications and Input/ Output Variables
Assumptions
•
•
Cooling Tower Specifications
Air exiting the tower is saturated with water
Rating (RT)
5000
vapor and is only characterized by its
Rating flow rate (GPM)
15300
enthalpy
Rating ambient wet bulb (F)
86
Rating ambient dry bulb (F)
122
Rating entering condenser water
105
Reduction of water flow rate by evaporation
is neglected in the energy balance.
•
Mass flow rate is calculated by considering
linear proportionality of mass flow rate of air
and motor speed.
temperature (F)
Fan diameter and speed (m, RPM)
8/152.6
Air flow rate (CFM)
776383
Inputs
•
•
•
•
•
Wet-bulb temperature
Cooling tower supply water temperature
Dry-bulb temperature
Mass flow rate of water
Cooling tower fan/motor speed
Outputs
•
•
Cooling tower return water temperature
Merkel’s Number
30
Cooling Tower Model
31
Pump Model
Mainly there are two mode of operation for these pumps:
Constant flow pump
Variable flow pump with a variable speed drive
To model a variable pump power following relationship is used:
𝑃𝑝𝑜𝑤𝑒𝑟 = 𝑃𝑀𝑃(𝐶1 + 𝑃𝐿𝑅(𝐶2 ) + 𝑃𝐿𝑅(𝐶3 )2 +𝑃𝐿𝑅(𝐶4 )3 )
Where,
PMP = pump motor power at rated condition, kW
C1, C2, C3 and C4 are pump performance coefficients
Also,
PLRi = pump part load ratio defined as follows:
𝑤𝑎𝑡𝑒𝑟 𝑓𝑙𝑜𝑤 𝑟𝑎𝑡𝑒 𝑉𝑖
𝐺𝑃𝑀
𝑃𝐿𝑅𝑖 =
=
𝑝𝑢𝑚𝑝 𝑐𝑎𝑝𝑎𝑐𝑖𝑡𝑦 𝑉𝑅
𝐺𝑃𝑀
32
Pump Model
Validation Graph
+ 5%Error Line
33
Solver Description
Qt,e
Tw,in,e
Tw,in,c
Ve
Vc
dTsh,e
34
Optimization
Optimization performed with two configurations:
Chiller Water Flow Optimization
Chiller Water Flow And Condenser Water Flow Optimization
Objective Function:
Minimize total power consumption i.e. compressor power and pump(s)
power combined.
35
Optimization
Chiller Water Flow Optimization:
Tw,in,c = 25 C and Tw,in,e = 14 C
Vc
Qe
0.4795 m3/s
10000 KW
Power
Ve
Total (KW) (m3/s)
2791.90
2494.66
2325.43
2226.70
2171.34
2145.79
2149.01
2177.04
2227.85
2300.43
2389.48
0.1419
0.1774
0.2129
0.2484
0.2839
0.3194
0.3548
0.3903
0.4258
0.4613
0.4968
COP
3.58
4.01
4.30
4.49
4.61
4.66
4.65
4.59
4.49
4.35
4.19
Vc
Qe
0.4795 m3/s
8000 KW
Power
Ve
Total (KW) (m3/s)
1768.81
1617.53
1535.14
1492.30
1476.36
1483.94
1512.07
1559.59
1623.07
1708.17
1809.82
0.1419
0.1774
0.2129
0.2484
0.2839
0.3194
0.3548
0.3903
0.4258
0.4613
0.4968
COP
4.52
4.95
5.21
5.36
5.42
5.39
5.29
5.13
4.93
4.68
4.42
Vc
Qe
0.4795 m3/s
6000 KW
Power
Ve
Total (KW) (m3/s)
1102.50
1032.45
997.69
988.83
998.06
1023.13
1065.53
1122.92
1197.94
1288.91
1397.89
0.1419
0.1774
0.2129
0.2484
0.2839
0.3194
0.3548
0.3903
0.4258
0.4613
0.4968
COP
5.44
5.81
6.01
6.07
6.01
5.86
5.63
5.34
5.01
4.66
4.29
Vc
Qe
0.4795 m3/s
4000 KW
Power
Ve
Total (KW) (m3/s)
649.15
626.16
622.12
633.21
657.75
695.26
746.65
811.97
892.40
988.98
1103.05
0.1419
0.1774
0.2129
0.2484
0.2839
0.3194
0.3548
0.3903
0.4258
0.4613
0.4968
COP
6.16
6.39
6.43
6.32
6.08
5.75
5.36
4.93
4.48
4.04
3.63
36
Optimization
Total Power (KW)
Chiller Water Flow And Condenser Water Flow Optimization:
Qe
Ve,opt
Vc,opt
= 10,000 kW
= 0.349 m3/s
= 0.408 m3/s
Tw,in,e = 14 C; Tw,in,c = 25 C
37
Optimization
Total Power (KW)
Chiller Water Flow And Condenser Water Flow Optimization:
Qe
Ve,opt
Vc,opt
= 8,000 kW
= 0.296 m3/s
= 0.355 m3/s
Tw,in,e = 14 C; Tw,in,c = 25 C
38
Optimization
Total Power (KW)
Chiller Water Flow And Condenser Water Flow Optimization:
Qe
Ve,opt
Vc,opt
= 6,000 kW
= 0.249 m3/s
= 0.332 m3/s
Tw,in,e = 14 C; Tw,in,c = 25 C
39
Optimization
Total Power (KW)
Chiller Water Flow And Condenser Water Flow Optimization:
Qe
= 4,000 kW
Ve,opt = 0.205 m3/s
Vc,opt = 0.251 m3/s
Tw,in,e = 14 C; Tw,in,c = 25 C
40
Optimization
Chiller Water Flow And Condenser Water Flow Optimization:
Tw,in,e = 14 C; Tw,in,c = 25 C
41
Optimization
Chiller Water Flow And Condenser Water Flow Optimization:
42
Conclusions
Variable Speed compressor provide savings of 30-40%
Variable speed pump for water circulation play an imperative role in
reducing overall power consumption of chiller plant.
Modeling of chiller components can be performed with limited
engineering information from manufactures.
43
Future Work
More rigorous compressor loss model
Transient model for the condenser and evaporator
Cooling tower Model
Variable Speed condenser pump
Investigate the effect of pressure drop and resistance from
refrigerant side
44
Q&A
45