- Email: [email protected]
Techno-economic analysis and optimization of the heat recovery of utility boiler flue gas
Applied Energy 112 (2013) 907–917
Contents lists available at SciVerse ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Techno-economic analysis and optimization of the heat recovery of utility boiler flue gas Gang Xu, Shengwei Huang, Yongping Yang ⇑, Ying Wu, Kai Zhang, Cheng Xu National Thermal Power Engineering & Technology Research Center, North China Electric Power University, Beijing 102206, China
h i g h l i g h t s Four typical flue gas heat recovery schemes are quantitatively analyzed. The analysis considers thermodynamic, heat transfer and hydrodynamics factors. Techno-economic analysis and optimization design are carried out. High-stage steam substitute scheme obtains better energy-saving effect. Large heat transfer area and high flue gas resistances weaken overall performance.
a r t i c l e
i n f o
Article history: Received 28 September 2012 Received in revised form 8 April 2013 Accepted 12 April 2013 Available online 10 May 2013 Keywords: Heat recovery Utility boiler Flue gas Techno-economic analysis Optimization
a b s t r a c t Coal-fired power plants in China consume nearly half of available coals, and the resulting CO2 emissions cover over 40% of total national emissions. Therefore, reducing the energy expenditure of coal-fired power plants is of great significance to China’s energy security and greenhouse gas reduction programs. For coal-fired power plants, the temperature of a boiler’s exhaust gas reaches 120–150 °C or even higher. The thermal energy of boiler’s exhaust accounts for approximately 3–8% of the total energy of fuel input. Given these factors, we conducted a techno-economic analysis and optimization design of the heat recovery system using boiler exhaust gas. This research is conformed to the principles of thermodynamic, heat transfer, and hydrodynamics. Based on the data from an existing 1000 MW typical power generation unit in China, four typical flue gas heat recovery schemes are quantitatively analyzed from the thermodynamics perspective. The impacts of flue gas heat recovery on net work output and standard coal consumption rate of various schemes are performed. Furthermore, the transfer area of heat recovery exchanger and the draft fan work increment due to the flue gas pressure drop are analyzed. Finally, a techno-economic analysis of the heat recovery schemes is conducted, and some recommendations on optimization design parameters are proposed, with full consideration of various factors such as the decrease on fuel cost due to energy conservation as well as the investment cost of heat recovery retrofitting. The results revealed that, high-stage steam substitute scheme of flue gas heat recovery in power plant can obtain higher energy-saving effects than that of low-stage steam substitute scheme. And the energy-saving benefit of flue gas heat recovery, especially with high-stage steam substitute scheme, is weakened by large heat exchange areas and flue gas pressure drop caused by lower temperature differentials. Therefore, the techno-economic performance of the flue gas heat recovery in power plant may not always increase with the increment of the recovered heat but have an optimum point. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction In China, coal fired power plants consume nearly half of available coals; meanwhile, the resulting CO2 emissions cover over 40% of total national emissions. Therefore, energy conservation of coal-fired power plants is of great significance to China’s energy security and greenhouse gas control programs. ⇑ Corresponding author. E-mail addresses: [email protected], [email protected] (Y. Yang). 0306-2619/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2013.04.048
For coal-fired power plants, the temperature of a boiler’s exhaust gas can reach 120–150 °C or even higher. Its thermal energy accounts for approximately 50–80% of a boiler’s thermal loss and 3–8% of the total energy input of the power plant [1]. Moreover, the total energy of exhaust gas equals the heat value of nearly 70 million tons of standard coals. In other words, if 50% of the exhaust energy can be recovered, 35 million tons of standard coal can be saved annually, which is equivalent to the coal consumption of 40 sets of 600 MW coal-fired power units or to the annual power generation of two sets of the Three Gorges Power Station. Suppose
908
G. Xu et al. / Applied Energy 112 (2013) 907–917
Nomenclature
Abbreviations GTI Gas Technology Institute DOE Department of Energy DEA deaerator NAR net annual revenue Eu Euler number Re Reynolds number RH Regenerative heater ATD average temperature differential LHV low heat value WHE waste heat exchanger O&M operation and maintenance CRF capital recovery factor Symbols m h D Q
mass enthalpy volume flow of flue gas heat transfer capacity
the price of standard coal is 120 USD per ton, approximately 4.2 billion USD in fuel cost can be saved yearly. Recovering exhaust energy can also help to reduce CO2 emissions by approximately 100 million tons, which brings huge economic and social benefits. The boiler’s exhaust energy can generally be used to heat condensed water, cold air, and hot water of heating network [2–6]. Comparatively, the utilization of exhaust energy in the heating of condensed water is given considerable attention worldwide [7– 10]. On one hand, energy utilization in the heating of condensed water is a relatively mature technology and easy to realize; on the other hand, it can save a large amount of extracted steam to increase unit efficiency and to reduce energy consumption dramatically. As an example, the German Schwarze Pumpe power plant with a set of 2 800 MW lignite generation unit implements a flue gas division system after the electrostatic precipitator and uses exhaust energy to heat condensed water. Furthermore, the German Cologne Niederaussem 1000 MW power unit adopts the flue gas division system and delivers part of the flue gas to heat condensed water along the bypass flue, as a result of which, the exhaust heat of flue gas is fully utilized, and the flue gas temperature is further
4P he,I h0
gri Pnet Etotal bs s bf Tex CostInv Costom Inv
work output increment steam enthalpy of extraction steam steam enthalpy of exhaust steam internal efficiency of low-pressure part net output of power unit net input of power unit coal conservation rate increased coal consumption rate exhaust gas temperature annual capital cost annual O&M cost total investment cost
Subscripts w condensed water d drainage water e extraction steam in inlet of heaters out outlet of heaters
decreased. Since 2000, the Gas Technology Institute (GTI) in the US has studied transport membrane condenser technology to recover water and latent heat in the exhaust gas. In addition, the GTI has also conducted a series of industrial tests and commercial projects under the US Department of Energy [11]. In China, the Shanghai Waigaoqiao No. 3 power plant utilizes condensed water in the entrance of the 7th low-pressure regenerative heater to retrieve the heat energy of flue gas; this system reduces the design temperature of the flue gas from 125 °C to 85 °C, which improves boiler efficiency by 2%-points and overall unit efficiency by 0.8–0.9%points [12]. However, only a few studies have focused on heat recovery from the flue gas of a utility boiler since most studies have focused on the introduction of existing projects. In addition, seldom studies give importance to the thorough analysis and optimization of the heat recovery system. In view of these, this paper conducts systematic studies on the heat recovery of a utility boilers’ flue gas based on a typical ultrasupercritical power unit in China. The goals of this study are the following: (1) to evaluate the energy-saving effect of the heat
Fig. 1. Schematic of flue gas heat recovery system within power plant.
909
G. Xu et al. / Applied Energy 112 (2013) 907–917
recovery of utility boiler flue gas, (2) to reveal the influence of main heat recovery schemes on energy-saving effects, (3) to disclose the relationship of the heat recovery scheme with heat transfer area and flue gas pressure drop, and (4) to perform a techno-economic analysis and optimization of various heat recovery schemes.
2. Research foundation 2.1. Characteristics of heat recovery from a utility boiler For the power unit, the boiler’s exhaust energy can be used to heat condensed water, air, and heating water in the cogeneration unit. Among these applications, the easiest and most effective way to utilize exhaust energy is to heat condensed water. In a modern steam power plant, a large amount of steam is extracted to heat the condensed water. The heat will increase the temperature of condensed water; however, the extraction steam will compromise work output effectiveness. If the exhaust energy of the flue gas can be utilized to heat the condensed water and the extraction steam is reduced simultaneously, the saved extraction steam can also be utilized for more power output. Utilization of the extraction steam will raise work output and overall efficiency. Fig. 1 illustrates a typical flue gas heat recovery system in a steam-powered generation unit. In this figure, the waste heat of the boiler replaces part of the seventh steam extraction to heat the condensed water [13]. It is not difficult to find that the system with exhaust energy utilization has many special restrictions. In other words, the system is ‘‘restricted’’: (1) Flue gas inlet temperature in the exhaust heat exchanger is usually fixed. In addition, outlet temperature is restricted by material resistance to corrosion [14]. (2) Exhaust energy utilization is affected not only by flue gas temperatures and compositions in the boiler side, but also by extraction conditions in the turbine side.
IP T
HP T
Table 1 Overall performance of the base case. Items
Unit
Value
Low heat rate of coal (LHV) Coal input rate Total energy of coal input (LHV)
MJ/kg, ar kg/h MW
21.13 395,560 2321.7
Steam/water cycle performance Existing steam turbine generator output Total auxiliary power Net output
MW MW MW
1000 57.60 942.4
Overall plant performance Net efficiency Net coal consumption rate
% g/kW h
43.07 302.62
In view of the above restrictions, research on exhaust energy utilization of flue gas should not only analyze the heat transfer and flow characteristics of the exhaust heat exchanger but also investigate the flue gas working conditions and steam/water situations in the turbine side. Furthermore, better comprehensive assessments and optimization results can be achieved if disciplinary knowledge such as thermodynamics, heat transfer mechanics, fluid mechanics, and technical economics is combined. 2.2. Description of a typical coal-fired power generation unit A typical coal-fired power unit with ultra-supercritical parameters is selected as the research target in this paper. The plant uses bituminous coal, which contains 56.26% carbon, 3.79% hydrogen, 12.11% oxygen, 0.82% nitrogen, 0.17% sulfur, and 18.1% water, respectively. Under turbine heat acceptance (THA) working conditions, actual coal consumption rate is 395.56 t/h, boiler efficiency is 94.08%, and exhaust gas temperature is 131 °C. For the main steam, steam temperature and pressure are 600 °C and 26.25 MPa respectively with flow rate of 2707.3 t/h. For the reheated steam, steam temperature and pressure are 600 °C and 5 MPa. Additionally, the
LP T
LP T
DEA
RH1
RH2
RH3
COND
RH5
RH6
RH7
RH8
Fig. 2. Steam/water cycle of the 1000 MW supercritical power unit.
910
G. Xu et al. / Applied Energy 112 (2013) 907–917
exhaust parameters of the low pressure part of steam turbine are 0.00576 MPa and 35.42 °C. The steam/water system is shown in Fig. 2. This unit adopts 8-stage extraction processes, including 3-stage high-pressure heaters, 5-stage low-pressure heaters, and one deaerator. Here, the deaerator can also be considered as one of low-pressure heater. The overall performance of the unit is summarized in Table 1. The thermal parameters of the regenerative system are listed in Table 2. 2.3. Typical methods of exhaust energy utilization As we have mentioned before, the flue gas exhaust energy can replace part of the extraction steam in heating the condensed water to increase work output. Based on the location of condensed water, many cases of exhaust energy utilization can be implemented. Several typical cases are illustrated in Fig. 3, and are presented as follows: Case 1: The exhaust heat exchanger and 7th-stage regenerative heater are installed in parallel (as illustrated in Fig. 3a). Part of the condensed water at the entrance of the 7th-stage regenera-
tive heater will enter the low-temperature economizer and return to the thermal system after absorbing a certain amount of energy. Afterward, the condensed water will converge with the main condensed water at the exit of the 7th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 7th-stage regenerative heater. Case 2: The exhaust heat exchanger and 8th-stage regenerative heater are installed in parallel (Fig. 3b). Part of the condensed water at the entrance of the 8th-stage regenerative heater will the enter low-temperature economizer and return to the thermal system after absorbing a certain amount of energy. Afterward, the condensed water will converge with the main condensed water at the exit of the 8th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 8th-stage regenerative heater. That is, this case is a kind of low-stage steam substitute scheme. Case 3: The exhaust heat exchanger, 7th-stage regenerative heater, together with the 8th-stage regenerative heater are installed in parallel (Fig. 3c). Part of the condensed water at the entrance of the 8th-stage regenerative heater will the enter low-temperature economizer and return to the thermal system
Table 2 Main parameters of the regenerative system. Items
Regenerative system parameters
Extracted steam pressure (MPa) Extracted steam temperature (°C), Extracted steam flow rate (kg/s) Inlet feedwater temperature (°C) Outlet feedwater temperature (°C)
RH1
RH2
RH3
DEA
RH5
RH6
RH7
RH8
8.307 389.0 39.98 268.7 290.0
6.103 351.2 75.80 219.4 268.70
2.518 492.6 31.28 189.9 219.4
1.234 380.5 26.20 153.3 183.3
0.626 288.6 30.74 122.1 153.3
0.2556 192.1 35.40 83.3 122.1
0.0671 86.1 21.17 60.8 83.3
0.0257 68.7 20.96 38.9 60.8
RH6
tyy
RH7
RH7
tyy
RH8
Waste heat exchanger
Waste heat exchanger
tpy
tpy
(a) Waste heat exchanger parallel to the 7th-stage
(b) Waste heat exchanger parallel to the 8th-stage
regenerative heater
regenerative heater
tyy
RH7
RH8
RH6 tyy Waste heat exchanger
Waste heat exchanger tpy
tpy
(c) Waste heat exchanger parallel to the 7th and th
RH7
8 -stage regenerative heaters
(d) Waste heat exchanger Series to the 6th and 7thstage regenerative heaters
Fig. 3. Main schemes of flue gas heat recovery within power plant.
911
G. Xu et al. / Applied Energy 112 (2013) 907–917
after absorbing a certain amount of energy. Afterward, the condensed water will converge with the main condensed water at the exit of the 7th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 7th-stage and 8th-stage regenerative heaters. Case 4: The exhaust heat exchangers, 6th-stage regenerative heater, together with the 7th-stage regenerative heater are installed in series (Fig. 3d). A shunt valve is installed at the exit of the 7th-stage regenerative heater to regulate part or all of the condensed water that enters the low-temperature economizer. Additionally, a booster pump is needed to deliver the condensed water to the thermal system. Afterward, the condensed water will converge with the main condensed water at the entrance of the 6th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 6th-stage regenerative heater. That is, this case is a kind of high-stage steam substitute scheme. In all these cases, the temperature of the exhaust flue gas is assumed to be at 131 °C. Considering the coal ranks, materials of flue gas duct, acid corrosion points, and other factors, the minimum exhaust gas temperature is set at 90 °C. Combined with the practical situation, the four cases listed above represent several typical situations in the power unit. 3. Theoretical analysis of flue gas heat recovery system 3.1. Thermodynamic analysis 3.1.1. Heat equivalence of flue gas heat recovery process For a specific heater (Fig. 4), thermal equilibrium can be represented in the following form: 0
0
0
0
ðm0w;out hw;out m0w;in hw;in Þ ¼ m0e;i he;i þ m0d;in hd;in þ Q 0
m0d;out hd;out
ð1Þ
Suppose the energy of flue gas exhaust is ignored for the same heater, thermal equilibrium can be expressed as follows:
ðmw;out hw;out mw;in hw;in Þ ¼ me;i he;i þ md;in hd;in md;out hd;out
ð2Þ
where m, h, and Q are the mass, enthalpy, and absorbed exhaust heat of the flue gas; for the subscripts, w, d, and e are the condensed water, drainage water, and extraction steam; and in and out are the inlet and outlet condition of the heaters. The exhaust utilization parameter is expressed with single quotation marks, whereas the non-utilization parameter is shown without any quotation marks.
3.1.2. Calculation of the saved steam extraction The saved steam extraction Dm can be expressed as the following:
Dm ¼ me;i m0e;i
ð3Þ
According to the working conditions (m0w;out ¼ mw;out , 0 0 0 hw;out ¼ hw;out , m0d;in ¼ md;in , hd;in ¼ hd;in , he;i ¼ he;i ), mass conservation (m0w;out ¼ m0w;in þ m0e , mw,out = mw,in + me, m0d;out ¼ m0d;in þ m0e and md,out = md,in + me), and Eqs. (1)–(3), the relationship between the saved steam extraction Dm and exhaust gas temperature are obtained (Fig. 5). In Fig. 5, the recovered energy and substitute extraction steam Dm will augment with the drop of exhaust gas temperature; therefore, the substitute extraction steam Dm in different cases is similar to each other when the exhaust gas temperature remains constant. 3.1.3. Calculation of work output increment Saved steam extraction increases the work output of a turbine. The increase in work output DP can be expressed as the following:
DP ¼ Dm ðhe;i h0 Þ gri =3600
ð4Þ
where he,i and h0 denote the steam enthalpy of extraction steam and exhaust steam, respectively; gri is the internal efficiency of the lowpressure part of steam turbine. Based on the above analysis, the relationship between net increased output DP and exhaust gas temperature Tex in the each scheme is obtained (Fig. 6). Based on Fig. 6, work output increases with the decrease of exhaust gas temperature. However, the increase in work output under different cases greatly differs when exhaust gas temperature is constant. Compared with Case 2, the work output increase in Case 4 is four or five times larger. As seen in Eq. (4), DP varies with Dm, he,i, h0, and gri. In Fig. 5, when exhaust gas temperature remains constant, Dm in different cases only changes slightly and has little influence on exhaust enthalpy h0 and internal efficiency gri. Therefore, DP in different cases is mainly affected by extraction steam enthalpy he,i. The extraction steam enthalpy he,i varies in different cases; thus, DP in different cases vary dramatically. 3.2. Heat transfer analysis 3.2.1. Calculation of average temperature difference In this paper, the waste heat exchanger operates at low temperatures and has a low risk of overheating. Therefore, a counter-cur22 case 1 case 2 case 3 case 4
mw,out, hw,out md,in, hd,in
mw,in, hw,in RHi
Waste heat exchanger
md,out, hd,out
Saved steam extraction, kg/s
m e , he 20 18 16 14 12 10
Q
110
105
100
95
Exhaust gas temperature, ºC Fig. 4. Heat balance of a regenerative heater.
Fig. 5. Variation curves of saved steam extraction.
90
912
G. Xu et al. / Applied Energy 112 (2013) 907–917
12
Work output increment, MW
case 4 10
8
case 1
6
case 3 4
case 2 2 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 6. Variation trends of work output increment.
rent layout is selected to enlarge the temperature differential and reduce the heat transfer area. For the exhaust heat exchanger, the heat transfer equation is expressed as follows:
Q ¼ kADt m
Dt d Dt x lnðDtd =Dt x Þ
ð6Þ
where Dt d ¼ t01 t 002 and Dt x ¼ t 001 t02 , t 01 stand for the exhaust temperature without the heat exchanger; t02 and t 002 stand for the inlet and outlet temperature of condensed water in the heat exchanger; and t 001 stands for the exhaust flue gas temperature when the heat exchanger is installed. Based on the aforementioned analysis, the average temperature differential curve of the decrease in flue gas temperature is illustrated in Fig. 7. In Fig. 7, the temperature differential in all cases decreases with the drop of exhaust gas temperature. The energy-saving effect of
hy ¼ 0:2 C z C s ky Re0:65 Pr0:33 =d ðW=ðm2 KÞÞ f f
70 60
case 2
50
case 3
ha ¼ 0:023 C t C d C l ks ðws d=ms Þ Pr0:4 =d ðW=ðm2 KÞÞ
case 1
ð8Þ
where Ct refers to the temperature correction coefficient, and is approximately equivalent to 1; Cd refers to the passage shape correction coefficient, and is approximately equivalent to 1 for the circular tube; Cl refers to the tube length correction coefficient, and is approximately equivalent to 1; ws, ks , ms, and Pr refer to the inner tube velocity of inner tube condensed water, thermal coefficient, kinematic viscosity coefficient, and Prandtl number, respectively. In engineering applications, heat transfer calculation of a spiral finned tube can be corrected on the foundation of basic heat transfer calculations. Compared with light tubes, the thermal resistance of a flue gas outside a spiral finned tube can be corrected by the following equation:
Ry ¼
1 ððm2 KÞ=WÞ hy b1
ð9Þ
where b1 is the ratio of the external surface area to the internal surface area of thermal convection. Compared with light tubes, the thermal resistance of spiral finned tube materials can be corrected by the following equation:
d ððm2 KÞ=WÞ kb b2
ð10Þ
where b2 is the ratio of the external surface area to the internal surface area of thermal conduction. When waste heat is lengthways transferred lengthways to condensed water, thermal resistance can be calculated as follows:
Ra ¼ 40
ð7Þ
where ky , Re, and Pr independently represent the thermal conductivity, Reynolds number, and Prandtl number of the flue gas; Cz and Cs represent the tube correction coefficient and arrangement correction coefficient, respectively. The tube correction coefficient can be calculated as follows:When tube numbers z < 10, Cz = 0.91 + 0.0125 d2, when tube numbers z P 10; C z ¼ 1. For the heat transfer process wherein the condensed water lengthways passes lengthways against the inner wall of the heat pipe, the heat transfer coefficient in the condensed water side can be calculated using the following formula:
Rb ¼
80
Average temperature differential, ºC
3.2.2. Calculation of heat transfer coefficient Flue gas heats condensed water through light tubes; the related equation is expressed as follows [16]:
ð5Þ
where Q, k, and DP are the heat transfer capacity, overall heat transfer coefficient, and heat transfer area, respectively; DI is the average temperature differential of the heat exchanger. For the heat exchanger with a counter-current layout, Dtm can be expressed as follows [15]:
Dt m ¼
the high-stage steam substitute scheme is more effective (Section 3.1); however, its temperature differential of the substitute scheme is lower and decreases at a faster rate. Among the four cases, temperature differential is the highest and decreases at a faster rate in Case 4. When the exhaust gas temperature reaches 90 °C, average temperature differential is approximately 15 °C, which is close to the engineering limit [14].
1 ððm2 KÞ=WÞ ha
ð11Þ
30
The overall heat transfer coefficient is expressed as the following:
20
K¼ case 4
1 ðW=ðm2 KÞÞ Ry þ Rb þ Ra
ð12Þ
10 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 7. Variation trends of the average temperature differential.
3.2.3. Calculation of heat transfer area The heat transfer coefficient can be calculated by the aforementioned corrected equation. Finally, heat transfer areas can be calculated by the heat transfer formula:
913
G. Xu et al. / Applied Energy 112 (2013) 907–917
60000
1600
case 4 1400
50000
40000
30000
20000
case 1
10000
case 3 case 2
The increment of flue gas pressure drop, Pa
Heat exchange area, m2
case 4
1200 1000 800 600
case 1 case 3 case 2
400 200
110
105
100
95
90
110
Q ðm2 Þ K Dt m
100
95
90
Fig. 9. Variation curves of flue gas pressure drop.
Fig. 8. Variation curves of heat transfer area.
A¼
105
Exhaust gas temperature, ºC
Exhaust gas temperature, ºC
ð13Þ
In this paper, all cases use spiral finned tubes. According to Eqs. (9)–(13), we get the curve relationship between heat transfer area and exhaust gas temperature (Fig. 8). From Eq. (13), we conclude that when heat transfer capacity (or exhaust gas temperature) remains constant, heat transfer area is approximately inversely proportional to temperature differential. Thus, from Fig. 8, it is obvious that the highest substitute scheme in Case 4 has the largest heat transfer area and increases exponentially.
of flue gas, wy is the average velocity of flue gas, z is the total tube numbers along the direction of flue gas flow. According to the above analysis, the curve relationship between flue gas pressure drop and exhaust gas temperature is illustrated as follows: From Eqs. (14)–(16) and Fig. 9, we can conclude that with the decline of exhaust gas temperature, flue gas pressure drop in all cases will increase. When exhaust gas temperature reaches 90 °C (that is, the temperature reduction of the flue gas in the heat exchanger gets to 40 °C), the flue gas pressure drop of the heat exchanger can reach 1512 Pa, which is nearly three times of that in Case 1 (573 Pa) and 5 times of that in Case 2 (359 Pa).
3.3. Fluid mechanic analysis 3.3.1. Flow parameters of flue gas The added heat exchanger will increase the pressure drop of flue gas, leading to the increment of induced fan work. And the increment of induced fan work is approximately proportional to flue gas capacity and pressure drop. Flue gas capacity can be calculated as follows:
3.3.3. Fan work increment caused by the increment of flue gas pressure drop Based on flue gas capacity in the heat exchanger, the increment of flue gas pressure drop, and induced draft fan efficiency, the increase of fan power can be calculated as follows:
Nf ¼
o
D T in þ 273:15 273:15 3600
ð14Þ
In this formula, D is the actual volume flow of flue gas (m3/s) in the heat exchanger; D0 is the standard volume flow of flue gas (m3 n/h) of the boiler; Tin is the inlet flue gas temperature in the system (131 °C is selected for this paper). For the heat exchanger, high frequency finned tubes with inline positions are selected; for single row serrated finned tubes, the following equation is adopted [16–18]:
c d pf S1 Eu ¼ aReb d0 d0
ð15Þ p
where Re stands for Reynolds number, d0f stands for the ratio of finned pitch space to wing space, dS10 stands for the relative pitch of high frequency finned tubes, and a, b, c, and d are empirical data obtained from related design parameter diagrams. Here, the quantitative value of a, b, c, and d are selected as 1, 2, 3, and 4, respectively. 3.3.2. The increment of flue gas pressure drop due to waste heat recovery Flue gas pressure drop can be further calculated as follows:
DP ¼ Eu q
w2y
z
ð16Þ
ð17Þ
In this formula, DNf represents increment of fan work after adopting the waste heat recovery (kW), gf refers to induced draft fan efficiency (%). And D is flue gas flow (m3 n/s). Eq. (17) reveals that the energy consumption of a booster fan is affected by flue gas flow D, the increment of flue gas pressure drop
2200
The increment of fan work, kw
D¼
D DP 1000 gf
case 4
2000 1800 1600 1400 1200 1000 800
case 1 case 3 case 2
600 400 200 110
105
100
95
Exhaust gas temperature, ºC In the formula, DP refers to the increment of flue gas pressure drop (Pa), Eu represents characteristics of tube resistance, q is the density
Fig. 10. Variation curves of increased fan power.
90
G. Xu et al. / Applied Energy 112 (2013) 907–917
DP, and fan efficiency gf. Among these factors, flue gas flow is mainly influenced by coal ranks and unit loads. Fan efficiency is usually dependent on fan shape selection, design, and operation adjustment. From the above formula, when flue gas capacity D and fan efficiency gf remain constant, the increment of fan work DNf has an approximately linear relationship with the increment of flue gas pressure drop DP. Fig. 10 shows the variation curve between the increment of fan work and flue gas temperature when the waste heat exchanger is added. We can also conclude that the electric consumption curve is proximate to flue gas pressure drop curve (Fig. 9). 4. Discussion 4.1. Energy conservation effect of heat recovery system 4.1.1. Brief introduction of evaluation criteria In the electricity industry, standard coal consumption rates are commonly applied to evaluate the thermal performance of coalfired power generation units. Standard coal consumption rate bs is defined as the following: s
b ¼
Etotal 3600=Hsu Etotal ¼ 122:8 ðg=kW hÞ Pnet P net
ð18Þ
In Eq. (18), Etotal refers to the total energy input per unit time, and its quantitative value is approximately equivalent to the total heat input value of fossil fuel per unit time; Pnet refers to net power output, which removes auxiliary power from gross power output. Etotal and Pnet should have the same unit of measure such as kW, MW, GW, or kW h. Number 3600 refers to 3600 kJ/kW h; Hsu , with a unit of kJ/g, refers to the heat value of standard coal (usually LHV in China), its value is 29.308 kJ/g or 7000 kcal/kg; the unit of standard coal consumption rate bs is g/kW h, representing standard coal consumption in grams per kW h of electricity output. Taking the net efficiency gnet = Pnet/Etotal into Eq. (18), it is easy to find that bs = 122.8/gnet. In other words, standard coal consumption rate has a direct relationship with net efficiency. Thus, bs reflects the thermal and economic performance of a power unit in a more effective manner. The selected base power unit mentioned above has a net power output of 942.4 MW and input fuel value Etotal of 2321.7 MW. According to Eq. (18), net efficiency and standard coal consumption rate is 43.07% and 302.62 g/kW h, respectively. When the waste heat of flue gas is recovered, part of the steam extracted for condensed water heating is saved, which will in turn increase the work output of the power generation unit. Taking work output increase DP into account, we obtain the standard coal consumption rate after waste heat recovery: s
bHR ¼ 122:8
Etotal ðg=kW hÞ Pnet þ DP
ð19Þ
Therefore, the coal conservation rate Dbs is deduced as the following: s
s
Db ¼ b
s bHR
DP ¼b ðg=kW hÞ Pnet þ DP s
ð20Þ
From Eqs. (19) to (20), we determine that standard coal consumption rate bs is a constant value because net output power Pnet and net efficiency gnet remain unchanged. As a result, coal conservation rate Dbs has a corresponding relation with increasing work output DP.As for the waste heat recovery system, the coal conservation rate Dbs can be easily calculated. Moreover, the coal conservation rate can directly reflect the energy-saving effect of a power unit after waste heat recovery. Thus, the coal conservation rate is selected
as the main evaluation criteria in the following chapters for the energy-saving effects of flue gas waste heat utilization. 4.1.2. Gross coal conservation rate From Eqs. (19) to (20) and in accordance with Table 1 and Fig. 6, we obtain the relationship between gross coal conservation rate and exhaust gas temperature. Fig. 11 reveals the following: (1) gross coal conservation rate Dbs has a similar variation curve with net output power increment DP under exhaust gas temperature Tex. The similarity is attributed to the interrelationship of Dbs and DP, as illustrated in Eq. (20); (2) the replaced steam in Case 4 has the highest stage number and pressure, which leads to the highest net work output increment DP. As a result, the gross coal conservation rate Dbs in Case 4 is larger than that in other cases; (3) the maximum value of gross coal conservation rate Dbs in Case 4 reaches up to 3.27 g/kW h, which is 1.79 times of that in Case 1 and 3.18 times of that in Case 2. The results show the remarkable energy-saving effects of Case 4. 4.1.3. Additional coal consumption rate due to the increment of flue gas pressure drop In practical engineering operations, flue gas pressure drop is increased by installing a heat exchanger in the tail flue, which leads to fan work increment. As a consequence, the coal consumption rate of a power unit is also increased. Thus, additional coal consumption rates and consequent effects caused by the increment of flue gas pressure drop in each scheme should be taken into consideration. Combining Eqs. (18)–(20) and considering the fan work caused s by flue gas pressure drop, the increased coal consumption rate bf is calculated as follows: s
s
bf ¼ b
Nf ðg=kW hÞ Pnet þ Nf
ð21Þ
Therefore, the variation curve of increased coal consumption caused by fan work is deduced (Fig. 12). 4.1.4. Net coal conservation rate After comprehensive consideration of the gross coal conservation rate due to waste heat utilization (Fig. 11) and the additional coal consumption rate caused by the increment of flue gas pressure drop (Fig. 12), the variation rule of net coal conservation rate is obtained (Fig. 13). From Figs. 11–13, we can find that: (1) with the drop of exhaust gas temperature Tex, gross coal conservation rate, the additional coal consumption rate and net coal conservation rate of all cases
Gross coal conservation rate, g/KWh
914
3.5 case 4
3.0 2.5 2.0
case 1
1.5
case 3 case 2
1.0 0.5 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 11. Variation curves of gross coal conservation rate.
915
0.7
case 4 0.6 0.5 0.4 0.3
case 1 0.2
case 3 case 2
0.1 110
105
100
95
90
Annual coal-saving revenue, Million USD
Additional coal consumption rate, g/kWh
G. Xu et al. / Applied Energy 112 (2013) 907–917
1.8
case 4
1.6 1.4 1.2 1.0
case 1
0.8
case 3
0.6
case 2
0.4 0.2 110
Exhaust gas temperature, ºC
105
100
95
90
85
Exhaust gas temperature, ºC Fig. 12. Variation curves of additional coal consumption rate. Fig. 14. Variation curves of annual coal-saving revenue.
case 4
2.5
2.0 case 1
1.5
case 3
1.0 case 2
shown in Fig. 14. The basic economic assumptions employed here include: (1) The assumed coal price, 120 USD per ton standard coal (750 CNY per ton standard coal), was the average cost to China electric generators in 2012. (2) The exchange rate is set as 6.25 CNY/USD. (3) The annual utilization hours of the unit are set as 5000 h. From Fig. 14, we can see that the annual coal-saving revenue and the net coal conservation rate have similar variation curves. The annual amount and revenue of coal saving in Case 4 can reach nearly 13,000 tons and 1.56 million USD per year with minimal exhaust gas temperatures (90 °C).
0.5
110
105
100
95
90
Exhaust gas temperature, ºC Fig. 13. Variation curves of net coal conservation rate.
will increase; (2) the gross coal conservation rate, additional coal consumption and net coal-saving consumption of the high-stage steam substitute scheme. (Case 4) is much larger than other schemes; (3) for gross coal conservation rate curve, all of the four cases increases linearly with the decline of Tex. Case 4 is the highest case. As shown in Fig. 11, when Tex equals to 90 °C, the gross coal conservation rate of Case 4 can reach 3.27 g/kW h, which is nearly 1.79 times of that in Case 1 and 3.15 times of that in Case 2. As shown in Fig. 11, (4) for additional coal consumption rate curve, all of the four cases increases rapidly with the decline of Tex. Among them, Case 4 is also the highest one and exponential growing with the increment of Tex. As shown in Fig. 12, when Tex equals to 90 °C, the additional coal conservation rate of Case 4 can reach 0.657 g/ kW h, which is nearly 2.63 times of that in Case 1 and 4.63 times of that in Case 2, and (5) for net coal conservation rate curve, all of the four cases increases with the decline of Tex. As shown in Fig. 13, Case 4 still ranks first among the cases. But its curve is obviously more flat than other cases. When Tex equals to 90 °C, the net coal conservation rate of Case 4 can reach 0.657 g/kW h, which is nearly 1.65 times of that in Case 1 and 2.96 times of that in Case 2.
4.2.2. Analysis of annual capital cost caused by waste heat recovery According to data of similar domestic retrofitting projects of utility boilers in China, the capital investment of a unit heat transfer area is approximately 120 USD/m2 [12]. Therefore, according to the heat transfer areas of various cases (Fig. 8), we calculated the total investment cost (Fig. 15). To evaluate the costs and benefits of the engineering project in different periods, we convert capital cost using equivalent conversion. According to the model of equivalent payment calculation, we introduce the capital recovery factor (CRF) concept [2,19,20]:
CRF ¼ ½i ð1 þ iÞn =½ð1 þ iÞn 1
ð22Þ
7
case 4 Total investment, Million USD
Net coal conservation rate, g/kWh
3.0
6 5 4 3
case 1 2
case 3 case 2
1
4.2. Techno-economic analysis of heat recovery system 110
4.2.1. Annual coal-saving revenue due to waste heat recovery Based on the above analysis of net coal conservation rate, the annual coal-saving revenue of various cases can be obtained, as
105
100
95
90
Exhaust gas temperature, ºC Fig. 15. Variation curves of total investment of heat recovery system.
G. Xu et al. / Applied Energy 112 (2013) 907–917
Cost Inv ¼ CRF Inv
ð23Þ
where ‘Inv’ refers to the total investment capital. Fig. 16 illustrates the variation curve of annual capital cost in the four cases. The annual capital cost has a similar variation curve with the total investment (Fig. 15). With the drop of exhaust gas temperature Tex, both total investment and annual capital cost increase in all cases. However, total investment and annual capital cost in Case 4 still ranks first and increase fastest with decline of exhaust gas temperature (Figs. 15 and 16). The reason mainly lies in that the heat exchanger of Case 4 has the lowest temperature differential and the largest heat transfer area (Fig. 8). As shown in Fig. 16, when exhaust gas temperature Tex reaches 90 °C, the annual capital cost in Case 4 reaches the maximum, which is nearly 2.64 times of that in Case 1 and 4.22 times of that in Case 2. 4.2.3. Analysis of annual operation and maintenance (O&M) cost due to waste heat recovery To simplify O&M cost calculation, we assumed that annual O&M cost is proportional to the total investment cost (Inv) with proportionality coefficient aOM, i.e., CostOM = aOM Inv. According to related literature [19], aOM can be set as 4%, that is, the O&M costs are fixed at 4% of the total investment cost per year. Fig. 17 illustrates the variation curves of annual O&M in the four cases. It can be seen that, the variation curves of CostOM in Fig. 17 correspond to that of total capital requirement in Fig. 15. 4.2.4. Comprehensive economic analysis and optimization Based on the aforementioned analysis, we obtain the net annual revenue (NAR) as follows:
NAR ¼ Incomecoal Cap Cost OM
ð24Þ
Here CostOM represents the O&M cost of the waste recovery system. Fig. 18 illustrates the variation curves of NAR for the heat recovery system. As seen from this figure, Case 4 still ranks first among the cases. However, the NAR of Case 4 will increase to an optimal point before decreasing as exhaust gas temperature decreases. This is owing to the fact that, with the drop of exhaust gas temperatures, coal-saving rate and revenue increase greatly (Figs. 13 and 14), but on the other hand, heat exchange areas and capital cost
Annual Capital Cost, Million USD
0.7
case 4 0.6 0.5 0.4 0.3
case 1 0.2
case 3 case 2
0.1 0.0 110
105
100
95
Exhaust gas temperature, ºC Fig. 16. Variation curves of annual capital cost.
90
0.30
Annual O&M Cost, Million USD
where i refers to the fraction interest rate per year, and n refers to the number of years that the capital has been borrowed over a fixed rate of interest. In this paper, i = 0.08 and n = 20. Therefore, the annual capital cost (CostInv) can be calculated as follows:
case 4 0.25
0.20
0.15
case 1
0.10
case 3 case 2
0.05
0.00
110
105
100
95
90
Exhaust gas temperature, ºC Fig. 17. Variation curves of annual O&M cost.
1.0
Net annual revenue, Million USD
916
optimal
0.9 0.8 0.7
case 4
0.6
case 1
0.5
case 3 0.4
case 2
0.3 0.2 0.1 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 18. Variation curves of net annual revenue.
simultaneously increase on a large scale (Figs. 9 and 16), which counteracts part of coal-saving benefits. Finally, when temperature decrease reaches a certain value, the increasing capital cost will exceed the coal-saving benefits. As a consequence, the comprehensive performance curve will increase rather than decrease with variation of exhaust gas temperature. The optimal exhaust gas temperature is 98 °C, which enables the NAR to reach 0.889 million USD. Therefore, there is an interesting result. That is, from the techno-economic perspective, the performance of the high-stage steam substitute scheme of flue gas heat recovery system will not always increase with the increment of the recovered heat (the drop of exhaust gas temperature). Instead, there is an optimum point. Fig. 19 illustrates the variation rule of the optimal exhaust gas temperature, annual capital cost, annual O&M cost, annual coalsaving revenue and net annual revenue (NAR) in the best scheme (Case 4) when the coal price varies. As seen in Fig. 19, with the increase of coal price, the optimal exhaust flue gas temperature will gradually decrease and annual coal-saving revenue will increase dramatically. Meanwhile, the annual capital and O&M cost increases gently while the net annual revenue increases greatly. This is due to that with the increase of coal price, the lower exhaust gas temperature is suitable for the waste heat recovery system, as a consequence of which, more waste heat recovery and
104
Annual coal-saving revenue Optimal exhaust gas temperature
2.0
1.6
102 Net annual revenue
1.2 100 0.8 98 Annual capital cost
96
0.4
Annual O&M cost
Economic parameters, Million USD
Optimal exhaust gas temperature, ºC
G. Xu et al. / Applied Energy 112 (2013) 907–917
0.0 80
100
120
140
160
Standard coal price, USD/ton Fig. 19. Sensitivity analysis with standard coal price.
more energy-saving effects can be achieved. Meanwhile, the further decreased exhaust temperature of flue gas will lead to more heat surface, resulting in more annual capital and O&M cost. However, the increment of investment and O&M cost is obviously much less than that of annual coal-saving revenue, therefore, leads to the obvious increment of the net annual revenues. Thus, when the coal price is comparatively high, it is proper to adopt the design with more heat surfaces to recover more waste heat, as a consequence of which, the economic benefits caused by waste heat recover will also increase dramatically. 5. Conclusion We conducted a techno-economic analysis and optimization design of a heat recovery system using boiler exhaust gas. This study is based on the principles of thermodynamic engineering, heat transfer, and hydrodynamics together with data from an existing typical 1000 MW power generation unit in China. From the results of this study, important conclusions are obtained and a few interesting integration measures are proposed. (1) The effect of flue gas exhaust energy is affected not only by exhaust gas temperature in the boiler side but also by the integrated solution in the steam/water side. The grade of substitute extraction steam has a determinative influence on the energy-saving effect. (2) High-stage steam substitute scheme of flue gas heat recovery in power plant can obtain higher energy-saving effects than that of low-stage steam substitute scheme. (3) The energy-saving benefit, especially with high-stage steam substitute scheme, is weakened by large heat exchange areas and flue gas pressure drop caused by lower temperature differentials. (4) Techno-economic analysis reveals that the economic performance of the high-stage steam substitute scheme will not always increase with the increment of the recovered heat and have an optimum point. When the coal price is comparatively high, it is proper to adopt the design with more heat surfaces to recover more waste heat, as a consequence of which, the economic benefits caused by waste heat recover will also increase dramatically.
917
(5) Based on the data from an existing 1000 MW typical power generation unit in China, the annual amount and revenue of coal saving in an optimal high-stage steam substitute scheme (Case 4) can reach nearly 13,000 tons and 1.56 million USD per year with minimal exhaust gas temperatures (90 °C). However, the net annual revenue of the optimal high-stage steam substitute scheme (Case 4) can reach highest value (0.889 million USD) when the exhaust gas temperature is 98 °C.
Acknowledgements This paper is supported by the National Major Fundamental Research Program of China (2011CB710706), the National Nature Science Fund of China (U1261210, 51006034) and the 111 Project (B12034).
References [1] Liu ZY. Annual report on the development of China’s electric power industry. Beijing: Electricity Council; 2011 [in Chinese]. [2] Westerlund Lars, Hermansson Roger, Fagerström Jonathan. Flue gas purification and heat recovery: a biomass fired boiler supplied with an open absorption system. Appl Energy 2012;96:444–50. [3] Zhang JH, Zhou YL, Li Y, et al. Generalized predictive control applied in waste heat recovery power plants. Appl Energy 2013;102:320–6. [4] Blarke Morten B. Towards an intermittency-friendly energy system: comparing electric boilers and heat pumps in distributed cogeneration. Appl Energy 2012;91:349–65. [5] Chen Q, Finney Karen, Li HN, et al. Condensing boiler applications in the process industry. Appl Energy 2012;89:30–6. [6] Pandiyarajan V, Chinna Pandian M, Malan E, et al. Experimental investigation on heat recovery from diesel engine exhaust using finned shell and tube heat exchanger and thermal storage system. Appl Energy 2011;88:77–87. [7] Espatolero Sergio, Cortés Cristóbal, Romeo Luis M. Optimization of boiler coldend and integration with the steam cycle in supercritical units. Appl Energy 2010;87:1651–60. [8] Jeong K, Kessen MJ, Bilirgen H. Analytical modeling of water condensation in condensing heat exchanger. Int J Heat Mass Transfer 2010;53:2361–8. [9] Jannelli E, Minutillo M. Simulation of the flue gas cleaning system of an RDF incineration power plant. Waste Manage 2007;27:684–90. [10] Wang CJ, He BS, Sun SY, et al. Application of a low pressure economizer for waste heat recovery from the exhaust flue gas in a 600 MW power plant. Energy 2012 [online 20 February 2012]. [11] Wang DX, Bao A, Kunc W, et al. Coal power plant flue gas waste heat and water recovery. Appl Energy 2012;91:341–8. [12] Zhao ZJ, Feng WZ, Zhang L, et al. Theoretical analysis and engineering practice of heat recovery form exhaust gas of power boilers. J Power Eng 2009;29:994–7 [in Chinese]. [13] Lin WC. Energy-saving theory of power plant system. Xi’an: Xi’an Jiaotong University Press; 1994 [in Chinese]. [14] Bahadori A. Estimation of combustion flue gas acid dew point during heat recovery and efficiency gain. Appl Therm Eng 2011;31:1457–62. [15] Manassaldi JI, Mussati SF, Scenna NJ. Optimal synthesis and design of heat recovery steam generation (HRSG) via mathematical programming. Energy 2011;36:475–85. [16] Yang SM, Tao WQ. Heat transfer. Beijing: Higher Education Press; 2006 [in Chinese]. [17] Kiyoshi K, Kenichi O, Takaharu K. Heat transfer and pressure drop characteristics of finned tube banks in forced convection (comparison of the pressure drop characteristics between spiral fin and serrated fin). Heat Transfer – Asian Res 2004;33(7):431–44. [18] Weierman C. Correlations ease the selection of fin tubes. Oil Gas J 1976;74(6):94–100. [19] Kreutz T, Williams R, Consonni S, Chiesa P. Co-production of hydrogen, electricity, and CO2 from coal with commercially ready technology. Part B: economic analysis. Hydrogen Energy 2005;30(7):769–84. [20] Xu G, Jin HG, Yang YP, et al. A comprehensive techno-economic analysis method for power generation systems with CO2 capture. Int J Energy Res 2010;34(4):321–32.
Contents lists available at SciVerse ScienceDirect
Applied Energy journal homepage: www.elsevier.com/locate/apenergy
Techno-economic analysis and optimization of the heat recovery of utility boiler flue gas Gang Xu, Shengwei Huang, Yongping Yang ⇑, Ying Wu, Kai Zhang, Cheng Xu National Thermal Power Engineering & Technology Research Center, North China Electric Power University, Beijing 102206, China
h i g h l i g h t s Four typical flue gas heat recovery schemes are quantitatively analyzed. The analysis considers thermodynamic, heat transfer and hydrodynamics factors. Techno-economic analysis and optimization design are carried out. High-stage steam substitute scheme obtains better energy-saving effect. Large heat transfer area and high flue gas resistances weaken overall performance.
a r t i c l e
i n f o
Article history: Received 28 September 2012 Received in revised form 8 April 2013 Accepted 12 April 2013 Available online 10 May 2013 Keywords: Heat recovery Utility boiler Flue gas Techno-economic analysis Optimization
a b s t r a c t Coal-fired power plants in China consume nearly half of available coals, and the resulting CO2 emissions cover over 40% of total national emissions. Therefore, reducing the energy expenditure of coal-fired power plants is of great significance to China’s energy security and greenhouse gas reduction programs. For coal-fired power plants, the temperature of a boiler’s exhaust gas reaches 120–150 °C or even higher. The thermal energy of boiler’s exhaust accounts for approximately 3–8% of the total energy of fuel input. Given these factors, we conducted a techno-economic analysis and optimization design of the heat recovery system using boiler exhaust gas. This research is conformed to the principles of thermodynamic, heat transfer, and hydrodynamics. Based on the data from an existing 1000 MW typical power generation unit in China, four typical flue gas heat recovery schemes are quantitatively analyzed from the thermodynamics perspective. The impacts of flue gas heat recovery on net work output and standard coal consumption rate of various schemes are performed. Furthermore, the transfer area of heat recovery exchanger and the draft fan work increment due to the flue gas pressure drop are analyzed. Finally, a techno-economic analysis of the heat recovery schemes is conducted, and some recommendations on optimization design parameters are proposed, with full consideration of various factors such as the decrease on fuel cost due to energy conservation as well as the investment cost of heat recovery retrofitting. The results revealed that, high-stage steam substitute scheme of flue gas heat recovery in power plant can obtain higher energy-saving effects than that of low-stage steam substitute scheme. And the energy-saving benefit of flue gas heat recovery, especially with high-stage steam substitute scheme, is weakened by large heat exchange areas and flue gas pressure drop caused by lower temperature differentials. Therefore, the techno-economic performance of the flue gas heat recovery in power plant may not always increase with the increment of the recovered heat but have an optimum point. Ó 2013 Elsevier Ltd. All rights reserved.
1. Introduction In China, coal fired power plants consume nearly half of available coals; meanwhile, the resulting CO2 emissions cover over 40% of total national emissions. Therefore, energy conservation of coal-fired power plants is of great significance to China’s energy security and greenhouse gas control programs. ⇑ Corresponding author. E-mail addresses: [email protected], [email protected] (Y. Yang). 0306-2619/$ - see front matter Ó 2013 Elsevier Ltd. All rights reserved. http://dx.doi.org/10.1016/j.apenergy.2013.04.048
For coal-fired power plants, the temperature of a boiler’s exhaust gas can reach 120–150 °C or even higher. Its thermal energy accounts for approximately 50–80% of a boiler’s thermal loss and 3–8% of the total energy input of the power plant [1]. Moreover, the total energy of exhaust gas equals the heat value of nearly 70 million tons of standard coals. In other words, if 50% of the exhaust energy can be recovered, 35 million tons of standard coal can be saved annually, which is equivalent to the coal consumption of 40 sets of 600 MW coal-fired power units or to the annual power generation of two sets of the Three Gorges Power Station. Suppose
908
G. Xu et al. / Applied Energy 112 (2013) 907–917
Nomenclature
Abbreviations GTI Gas Technology Institute DOE Department of Energy DEA deaerator NAR net annual revenue Eu Euler number Re Reynolds number RH Regenerative heater ATD average temperature differential LHV low heat value WHE waste heat exchanger O&M operation and maintenance CRF capital recovery factor Symbols m h D Q
mass enthalpy volume flow of flue gas heat transfer capacity
the price of standard coal is 120 USD per ton, approximately 4.2 billion USD in fuel cost can be saved yearly. Recovering exhaust energy can also help to reduce CO2 emissions by approximately 100 million tons, which brings huge economic and social benefits. The boiler’s exhaust energy can generally be used to heat condensed water, cold air, and hot water of heating network [2–6]. Comparatively, the utilization of exhaust energy in the heating of condensed water is given considerable attention worldwide [7– 10]. On one hand, energy utilization in the heating of condensed water is a relatively mature technology and easy to realize; on the other hand, it can save a large amount of extracted steam to increase unit efficiency and to reduce energy consumption dramatically. As an example, the German Schwarze Pumpe power plant with a set of 2 800 MW lignite generation unit implements a flue gas division system after the electrostatic precipitator and uses exhaust energy to heat condensed water. Furthermore, the German Cologne Niederaussem 1000 MW power unit adopts the flue gas division system and delivers part of the flue gas to heat condensed water along the bypass flue, as a result of which, the exhaust heat of flue gas is fully utilized, and the flue gas temperature is further
4P he,I h0
gri Pnet Etotal bs s bf Tex CostInv Costom Inv
work output increment steam enthalpy of extraction steam steam enthalpy of exhaust steam internal efficiency of low-pressure part net output of power unit net input of power unit coal conservation rate increased coal consumption rate exhaust gas temperature annual capital cost annual O&M cost total investment cost
Subscripts w condensed water d drainage water e extraction steam in inlet of heaters out outlet of heaters
decreased. Since 2000, the Gas Technology Institute (GTI) in the US has studied transport membrane condenser technology to recover water and latent heat in the exhaust gas. In addition, the GTI has also conducted a series of industrial tests and commercial projects under the US Department of Energy [11]. In China, the Shanghai Waigaoqiao No. 3 power plant utilizes condensed water in the entrance of the 7th low-pressure regenerative heater to retrieve the heat energy of flue gas; this system reduces the design temperature of the flue gas from 125 °C to 85 °C, which improves boiler efficiency by 2%-points and overall unit efficiency by 0.8–0.9%points [12]. However, only a few studies have focused on heat recovery from the flue gas of a utility boiler since most studies have focused on the introduction of existing projects. In addition, seldom studies give importance to the thorough analysis and optimization of the heat recovery system. In view of these, this paper conducts systematic studies on the heat recovery of a utility boilers’ flue gas based on a typical ultrasupercritical power unit in China. The goals of this study are the following: (1) to evaluate the energy-saving effect of the heat
Fig. 1. Schematic of flue gas heat recovery system within power plant.
909
G. Xu et al. / Applied Energy 112 (2013) 907–917
recovery of utility boiler flue gas, (2) to reveal the influence of main heat recovery schemes on energy-saving effects, (3) to disclose the relationship of the heat recovery scheme with heat transfer area and flue gas pressure drop, and (4) to perform a techno-economic analysis and optimization of various heat recovery schemes.
2. Research foundation 2.1. Characteristics of heat recovery from a utility boiler For the power unit, the boiler’s exhaust energy can be used to heat condensed water, air, and heating water in the cogeneration unit. Among these applications, the easiest and most effective way to utilize exhaust energy is to heat condensed water. In a modern steam power plant, a large amount of steam is extracted to heat the condensed water. The heat will increase the temperature of condensed water; however, the extraction steam will compromise work output effectiveness. If the exhaust energy of the flue gas can be utilized to heat the condensed water and the extraction steam is reduced simultaneously, the saved extraction steam can also be utilized for more power output. Utilization of the extraction steam will raise work output and overall efficiency. Fig. 1 illustrates a typical flue gas heat recovery system in a steam-powered generation unit. In this figure, the waste heat of the boiler replaces part of the seventh steam extraction to heat the condensed water [13]. It is not difficult to find that the system with exhaust energy utilization has many special restrictions. In other words, the system is ‘‘restricted’’: (1) Flue gas inlet temperature in the exhaust heat exchanger is usually fixed. In addition, outlet temperature is restricted by material resistance to corrosion [14]. (2) Exhaust energy utilization is affected not only by flue gas temperatures and compositions in the boiler side, but also by extraction conditions in the turbine side.
IP T
HP T
Table 1 Overall performance of the base case. Items
Unit
Value
Low heat rate of coal (LHV) Coal input rate Total energy of coal input (LHV)
MJ/kg, ar kg/h MW
21.13 395,560 2321.7
Steam/water cycle performance Existing steam turbine generator output Total auxiliary power Net output
MW MW MW
1000 57.60 942.4
Overall plant performance Net efficiency Net coal consumption rate
% g/kW h
43.07 302.62
In view of the above restrictions, research on exhaust energy utilization of flue gas should not only analyze the heat transfer and flow characteristics of the exhaust heat exchanger but also investigate the flue gas working conditions and steam/water situations in the turbine side. Furthermore, better comprehensive assessments and optimization results can be achieved if disciplinary knowledge such as thermodynamics, heat transfer mechanics, fluid mechanics, and technical economics is combined. 2.2. Description of a typical coal-fired power generation unit A typical coal-fired power unit with ultra-supercritical parameters is selected as the research target in this paper. The plant uses bituminous coal, which contains 56.26% carbon, 3.79% hydrogen, 12.11% oxygen, 0.82% nitrogen, 0.17% sulfur, and 18.1% water, respectively. Under turbine heat acceptance (THA) working conditions, actual coal consumption rate is 395.56 t/h, boiler efficiency is 94.08%, and exhaust gas temperature is 131 °C. For the main steam, steam temperature and pressure are 600 °C and 26.25 MPa respectively with flow rate of 2707.3 t/h. For the reheated steam, steam temperature and pressure are 600 °C and 5 MPa. Additionally, the
LP T
LP T
DEA
RH1
RH2
RH3
COND
RH5
RH6
RH7
RH8
Fig. 2. Steam/water cycle of the 1000 MW supercritical power unit.
910
G. Xu et al. / Applied Energy 112 (2013) 907–917
exhaust parameters of the low pressure part of steam turbine are 0.00576 MPa and 35.42 °C. The steam/water system is shown in Fig. 2. This unit adopts 8-stage extraction processes, including 3-stage high-pressure heaters, 5-stage low-pressure heaters, and one deaerator. Here, the deaerator can also be considered as one of low-pressure heater. The overall performance of the unit is summarized in Table 1. The thermal parameters of the regenerative system are listed in Table 2. 2.3. Typical methods of exhaust energy utilization As we have mentioned before, the flue gas exhaust energy can replace part of the extraction steam in heating the condensed water to increase work output. Based on the location of condensed water, many cases of exhaust energy utilization can be implemented. Several typical cases are illustrated in Fig. 3, and are presented as follows: Case 1: The exhaust heat exchanger and 7th-stage regenerative heater are installed in parallel (as illustrated in Fig. 3a). Part of the condensed water at the entrance of the 7th-stage regenera-
tive heater will enter the low-temperature economizer and return to the thermal system after absorbing a certain amount of energy. Afterward, the condensed water will converge with the main condensed water at the exit of the 7th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 7th-stage regenerative heater. Case 2: The exhaust heat exchanger and 8th-stage regenerative heater are installed in parallel (Fig. 3b). Part of the condensed water at the entrance of the 8th-stage regenerative heater will the enter low-temperature economizer and return to the thermal system after absorbing a certain amount of energy. Afterward, the condensed water will converge with the main condensed water at the exit of the 8th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 8th-stage regenerative heater. That is, this case is a kind of low-stage steam substitute scheme. Case 3: The exhaust heat exchanger, 7th-stage regenerative heater, together with the 8th-stage regenerative heater are installed in parallel (Fig. 3c). Part of the condensed water at the entrance of the 8th-stage regenerative heater will the enter low-temperature economizer and return to the thermal system
Table 2 Main parameters of the regenerative system. Items
Regenerative system parameters
Extracted steam pressure (MPa) Extracted steam temperature (°C), Extracted steam flow rate (kg/s) Inlet feedwater temperature (°C) Outlet feedwater temperature (°C)
RH1
RH2
RH3
DEA
RH5
RH6
RH7
RH8
8.307 389.0 39.98 268.7 290.0
6.103 351.2 75.80 219.4 268.70
2.518 492.6 31.28 189.9 219.4
1.234 380.5 26.20 153.3 183.3
0.626 288.6 30.74 122.1 153.3
0.2556 192.1 35.40 83.3 122.1
0.0671 86.1 21.17 60.8 83.3
0.0257 68.7 20.96 38.9 60.8
RH6
tyy
RH7
RH7
tyy
RH8
Waste heat exchanger
Waste heat exchanger
tpy
tpy
(a) Waste heat exchanger parallel to the 7th-stage
(b) Waste heat exchanger parallel to the 8th-stage
regenerative heater
regenerative heater
tyy
RH7
RH8
RH6 tyy Waste heat exchanger
Waste heat exchanger tpy
tpy
(c) Waste heat exchanger parallel to the 7th and th
RH7
8 -stage regenerative heaters
(d) Waste heat exchanger Series to the 6th and 7thstage regenerative heaters
Fig. 3. Main schemes of flue gas heat recovery within power plant.
911
G. Xu et al. / Applied Energy 112 (2013) 907–917
after absorbing a certain amount of energy. Afterward, the condensed water will converge with the main condensed water at the exit of the 7th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 7th-stage and 8th-stage regenerative heaters. Case 4: The exhaust heat exchangers, 6th-stage regenerative heater, together with the 7th-stage regenerative heater are installed in series (Fig. 3d). A shunt valve is installed at the exit of the 7th-stage regenerative heater to regulate part or all of the condensed water that enters the low-temperature economizer. Additionally, a booster pump is needed to deliver the condensed water to the thermal system. Afterward, the condensed water will converge with the main condensed water at the entrance of the 6th-stage regenerative heater. At this moment, the flue gas exhaust heat can substitute for the steam extraction of the 6th-stage regenerative heater. That is, this case is a kind of high-stage steam substitute scheme. In all these cases, the temperature of the exhaust flue gas is assumed to be at 131 °C. Considering the coal ranks, materials of flue gas duct, acid corrosion points, and other factors, the minimum exhaust gas temperature is set at 90 °C. Combined with the practical situation, the four cases listed above represent several typical situations in the power unit. 3. Theoretical analysis of flue gas heat recovery system 3.1. Thermodynamic analysis 3.1.1. Heat equivalence of flue gas heat recovery process For a specific heater (Fig. 4), thermal equilibrium can be represented in the following form: 0
0
0
0
ðm0w;out hw;out m0w;in hw;in Þ ¼ m0e;i he;i þ m0d;in hd;in þ Q 0
m0d;out hd;out
ð1Þ
Suppose the energy of flue gas exhaust is ignored for the same heater, thermal equilibrium can be expressed as follows:
ðmw;out hw;out mw;in hw;in Þ ¼ me;i he;i þ md;in hd;in md;out hd;out
ð2Þ
where m, h, and Q are the mass, enthalpy, and absorbed exhaust heat of the flue gas; for the subscripts, w, d, and e are the condensed water, drainage water, and extraction steam; and in and out are the inlet and outlet condition of the heaters. The exhaust utilization parameter is expressed with single quotation marks, whereas the non-utilization parameter is shown without any quotation marks.
3.1.2. Calculation of the saved steam extraction The saved steam extraction Dm can be expressed as the following:
Dm ¼ me;i m0e;i
ð3Þ
According to the working conditions (m0w;out ¼ mw;out , 0 0 0 hw;out ¼ hw;out , m0d;in ¼ md;in , hd;in ¼ hd;in , he;i ¼ he;i ), mass conservation (m0w;out ¼ m0w;in þ m0e , mw,out = mw,in + me, m0d;out ¼ m0d;in þ m0e and md,out = md,in + me), and Eqs. (1)–(3), the relationship between the saved steam extraction Dm and exhaust gas temperature are obtained (Fig. 5). In Fig. 5, the recovered energy and substitute extraction steam Dm will augment with the drop of exhaust gas temperature; therefore, the substitute extraction steam Dm in different cases is similar to each other when the exhaust gas temperature remains constant. 3.1.3. Calculation of work output increment Saved steam extraction increases the work output of a turbine. The increase in work output DP can be expressed as the following:
DP ¼ Dm ðhe;i h0 Þ gri =3600
ð4Þ
where he,i and h0 denote the steam enthalpy of extraction steam and exhaust steam, respectively; gri is the internal efficiency of the lowpressure part of steam turbine. Based on the above analysis, the relationship between net increased output DP and exhaust gas temperature Tex in the each scheme is obtained (Fig. 6). Based on Fig. 6, work output increases with the decrease of exhaust gas temperature. However, the increase in work output under different cases greatly differs when exhaust gas temperature is constant. Compared with Case 2, the work output increase in Case 4 is four or five times larger. As seen in Eq. (4), DP varies with Dm, he,i, h0, and gri. In Fig. 5, when exhaust gas temperature remains constant, Dm in different cases only changes slightly and has little influence on exhaust enthalpy h0 and internal efficiency gri. Therefore, DP in different cases is mainly affected by extraction steam enthalpy he,i. The extraction steam enthalpy he,i varies in different cases; thus, DP in different cases vary dramatically. 3.2. Heat transfer analysis 3.2.1. Calculation of average temperature difference In this paper, the waste heat exchanger operates at low temperatures and has a low risk of overheating. Therefore, a counter-cur22 case 1 case 2 case 3 case 4
mw,out, hw,out md,in, hd,in
mw,in, hw,in RHi
Waste heat exchanger
md,out, hd,out
Saved steam extraction, kg/s
m e , he 20 18 16 14 12 10
Q
110
105
100
95
Exhaust gas temperature, ºC Fig. 4. Heat balance of a regenerative heater.
Fig. 5. Variation curves of saved steam extraction.
90
912
G. Xu et al. / Applied Energy 112 (2013) 907–917
12
Work output increment, MW
case 4 10
8
case 1
6
case 3 4
case 2 2 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 6. Variation trends of work output increment.
rent layout is selected to enlarge the temperature differential and reduce the heat transfer area. For the exhaust heat exchanger, the heat transfer equation is expressed as follows:
Q ¼ kADt m
Dt d Dt x lnðDtd =Dt x Þ
ð6Þ
where Dt d ¼ t01 t 002 and Dt x ¼ t 001 t02 , t 01 stand for the exhaust temperature without the heat exchanger; t02 and t 002 stand for the inlet and outlet temperature of condensed water in the heat exchanger; and t 001 stands for the exhaust flue gas temperature when the heat exchanger is installed. Based on the aforementioned analysis, the average temperature differential curve of the decrease in flue gas temperature is illustrated in Fig. 7. In Fig. 7, the temperature differential in all cases decreases with the drop of exhaust gas temperature. The energy-saving effect of
hy ¼ 0:2 C z C s ky Re0:65 Pr0:33 =d ðW=ðm2 KÞÞ f f
70 60
case 2
50
case 3
ha ¼ 0:023 C t C d C l ks ðws d=ms Þ Pr0:4 =d ðW=ðm2 KÞÞ
case 1
ð8Þ
where Ct refers to the temperature correction coefficient, and is approximately equivalent to 1; Cd refers to the passage shape correction coefficient, and is approximately equivalent to 1 for the circular tube; Cl refers to the tube length correction coefficient, and is approximately equivalent to 1; ws, ks , ms, and Pr refer to the inner tube velocity of inner tube condensed water, thermal coefficient, kinematic viscosity coefficient, and Prandtl number, respectively. In engineering applications, heat transfer calculation of a spiral finned tube can be corrected on the foundation of basic heat transfer calculations. Compared with light tubes, the thermal resistance of a flue gas outside a spiral finned tube can be corrected by the following equation:
Ry ¼
1 ððm2 KÞ=WÞ hy b1
ð9Þ
where b1 is the ratio of the external surface area to the internal surface area of thermal convection. Compared with light tubes, the thermal resistance of spiral finned tube materials can be corrected by the following equation:
d ððm2 KÞ=WÞ kb b2
ð10Þ
where b2 is the ratio of the external surface area to the internal surface area of thermal conduction. When waste heat is lengthways transferred lengthways to condensed water, thermal resistance can be calculated as follows:
Ra ¼ 40
ð7Þ
where ky , Re, and Pr independently represent the thermal conductivity, Reynolds number, and Prandtl number of the flue gas; Cz and Cs represent the tube correction coefficient and arrangement correction coefficient, respectively. The tube correction coefficient can be calculated as follows:When tube numbers z < 10, Cz = 0.91 + 0.0125 d2, when tube numbers z P 10; C z ¼ 1. For the heat transfer process wherein the condensed water lengthways passes lengthways against the inner wall of the heat pipe, the heat transfer coefficient in the condensed water side can be calculated using the following formula:
Rb ¼
80
Average temperature differential, ºC
3.2.2. Calculation of heat transfer coefficient Flue gas heats condensed water through light tubes; the related equation is expressed as follows [16]:
ð5Þ
where Q, k, and DP are the heat transfer capacity, overall heat transfer coefficient, and heat transfer area, respectively; DI is the average temperature differential of the heat exchanger. For the heat exchanger with a counter-current layout, Dtm can be expressed as follows [15]:
Dt m ¼
the high-stage steam substitute scheme is more effective (Section 3.1); however, its temperature differential of the substitute scheme is lower and decreases at a faster rate. Among the four cases, temperature differential is the highest and decreases at a faster rate in Case 4. When the exhaust gas temperature reaches 90 °C, average temperature differential is approximately 15 °C, which is close to the engineering limit [14].
1 ððm2 KÞ=WÞ ha
ð11Þ
30
The overall heat transfer coefficient is expressed as the following:
20
K¼ case 4
1 ðW=ðm2 KÞÞ Ry þ Rb þ Ra
ð12Þ
10 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 7. Variation trends of the average temperature differential.
3.2.3. Calculation of heat transfer area The heat transfer coefficient can be calculated by the aforementioned corrected equation. Finally, heat transfer areas can be calculated by the heat transfer formula:
913
G. Xu et al. / Applied Energy 112 (2013) 907–917
60000
1600
case 4 1400
50000
40000
30000
20000
case 1
10000
case 3 case 2
The increment of flue gas pressure drop, Pa
Heat exchange area, m2
case 4
1200 1000 800 600
case 1 case 3 case 2
400 200
110
105
100
95
90
110
Q ðm2 Þ K Dt m
100
95
90
Fig. 9. Variation curves of flue gas pressure drop.
Fig. 8. Variation curves of heat transfer area.
A¼
105
Exhaust gas temperature, ºC
Exhaust gas temperature, ºC
ð13Þ
In this paper, all cases use spiral finned tubes. According to Eqs. (9)–(13), we get the curve relationship between heat transfer area and exhaust gas temperature (Fig. 8). From Eq. (13), we conclude that when heat transfer capacity (or exhaust gas temperature) remains constant, heat transfer area is approximately inversely proportional to temperature differential. Thus, from Fig. 8, it is obvious that the highest substitute scheme in Case 4 has the largest heat transfer area and increases exponentially.
of flue gas, wy is the average velocity of flue gas, z is the total tube numbers along the direction of flue gas flow. According to the above analysis, the curve relationship between flue gas pressure drop and exhaust gas temperature is illustrated as follows: From Eqs. (14)–(16) and Fig. 9, we can conclude that with the decline of exhaust gas temperature, flue gas pressure drop in all cases will increase. When exhaust gas temperature reaches 90 °C (that is, the temperature reduction of the flue gas in the heat exchanger gets to 40 °C), the flue gas pressure drop of the heat exchanger can reach 1512 Pa, which is nearly three times of that in Case 1 (573 Pa) and 5 times of that in Case 2 (359 Pa).
3.3. Fluid mechanic analysis 3.3.1. Flow parameters of flue gas The added heat exchanger will increase the pressure drop of flue gas, leading to the increment of induced fan work. And the increment of induced fan work is approximately proportional to flue gas capacity and pressure drop. Flue gas capacity can be calculated as follows:
3.3.3. Fan work increment caused by the increment of flue gas pressure drop Based on flue gas capacity in the heat exchanger, the increment of flue gas pressure drop, and induced draft fan efficiency, the increase of fan power can be calculated as follows:
Nf ¼
o
D T in þ 273:15 273:15 3600
ð14Þ
In this formula, D is the actual volume flow of flue gas (m3/s) in the heat exchanger; D0 is the standard volume flow of flue gas (m3 n/h) of the boiler; Tin is the inlet flue gas temperature in the system (131 °C is selected for this paper). For the heat exchanger, high frequency finned tubes with inline positions are selected; for single row serrated finned tubes, the following equation is adopted [16–18]:
c d pf S1 Eu ¼ aReb d0 d0
ð15Þ p
where Re stands for Reynolds number, d0f stands for the ratio of finned pitch space to wing space, dS10 stands for the relative pitch of high frequency finned tubes, and a, b, c, and d are empirical data obtained from related design parameter diagrams. Here, the quantitative value of a, b, c, and d are selected as 1, 2, 3, and 4, respectively. 3.3.2. The increment of flue gas pressure drop due to waste heat recovery Flue gas pressure drop can be further calculated as follows:
DP ¼ Eu q
w2y
z
ð16Þ
ð17Þ
In this formula, DNf represents increment of fan work after adopting the waste heat recovery (kW), gf refers to induced draft fan efficiency (%). And D is flue gas flow (m3 n/s). Eq. (17) reveals that the energy consumption of a booster fan is affected by flue gas flow D, the increment of flue gas pressure drop
2200
The increment of fan work, kw
D¼
D DP 1000 gf
case 4
2000 1800 1600 1400 1200 1000 800
case 1 case 3 case 2
600 400 200 110
105
100
95
Exhaust gas temperature, ºC In the formula, DP refers to the increment of flue gas pressure drop (Pa), Eu represents characteristics of tube resistance, q is the density
Fig. 10. Variation curves of increased fan power.
90
G. Xu et al. / Applied Energy 112 (2013) 907–917
DP, and fan efficiency gf. Among these factors, flue gas flow is mainly influenced by coal ranks and unit loads. Fan efficiency is usually dependent on fan shape selection, design, and operation adjustment. From the above formula, when flue gas capacity D and fan efficiency gf remain constant, the increment of fan work DNf has an approximately linear relationship with the increment of flue gas pressure drop DP. Fig. 10 shows the variation curve between the increment of fan work and flue gas temperature when the waste heat exchanger is added. We can also conclude that the electric consumption curve is proximate to flue gas pressure drop curve (Fig. 9). 4. Discussion 4.1. Energy conservation effect of heat recovery system 4.1.1. Brief introduction of evaluation criteria In the electricity industry, standard coal consumption rates are commonly applied to evaluate the thermal performance of coalfired power generation units. Standard coal consumption rate bs is defined as the following: s
b ¼
Etotal 3600=Hsu Etotal ¼ 122:8 ðg=kW hÞ Pnet P net
ð18Þ
In Eq. (18), Etotal refers to the total energy input per unit time, and its quantitative value is approximately equivalent to the total heat input value of fossil fuel per unit time; Pnet refers to net power output, which removes auxiliary power from gross power output. Etotal and Pnet should have the same unit of measure such as kW, MW, GW, or kW h. Number 3600 refers to 3600 kJ/kW h; Hsu , with a unit of kJ/g, refers to the heat value of standard coal (usually LHV in China), its value is 29.308 kJ/g or 7000 kcal/kg; the unit of standard coal consumption rate bs is g/kW h, representing standard coal consumption in grams per kW h of electricity output. Taking the net efficiency gnet = Pnet/Etotal into Eq. (18), it is easy to find that bs = 122.8/gnet. In other words, standard coal consumption rate has a direct relationship with net efficiency. Thus, bs reflects the thermal and economic performance of a power unit in a more effective manner. The selected base power unit mentioned above has a net power output of 942.4 MW and input fuel value Etotal of 2321.7 MW. According to Eq. (18), net efficiency and standard coal consumption rate is 43.07% and 302.62 g/kW h, respectively. When the waste heat of flue gas is recovered, part of the steam extracted for condensed water heating is saved, which will in turn increase the work output of the power generation unit. Taking work output increase DP into account, we obtain the standard coal consumption rate after waste heat recovery: s
bHR ¼ 122:8
Etotal ðg=kW hÞ Pnet þ DP
ð19Þ
Therefore, the coal conservation rate Dbs is deduced as the following: s
s
Db ¼ b
s bHR
DP ¼b ðg=kW hÞ Pnet þ DP s
ð20Þ
From Eqs. (19) to (20), we determine that standard coal consumption rate bs is a constant value because net output power Pnet and net efficiency gnet remain unchanged. As a result, coal conservation rate Dbs has a corresponding relation with increasing work output DP.As for the waste heat recovery system, the coal conservation rate Dbs can be easily calculated. Moreover, the coal conservation rate can directly reflect the energy-saving effect of a power unit after waste heat recovery. Thus, the coal conservation rate is selected
as the main evaluation criteria in the following chapters for the energy-saving effects of flue gas waste heat utilization. 4.1.2. Gross coal conservation rate From Eqs. (19) to (20) and in accordance with Table 1 and Fig. 6, we obtain the relationship between gross coal conservation rate and exhaust gas temperature. Fig. 11 reveals the following: (1) gross coal conservation rate Dbs has a similar variation curve with net output power increment DP under exhaust gas temperature Tex. The similarity is attributed to the interrelationship of Dbs and DP, as illustrated in Eq. (20); (2) the replaced steam in Case 4 has the highest stage number and pressure, which leads to the highest net work output increment DP. As a result, the gross coal conservation rate Dbs in Case 4 is larger than that in other cases; (3) the maximum value of gross coal conservation rate Dbs in Case 4 reaches up to 3.27 g/kW h, which is 1.79 times of that in Case 1 and 3.18 times of that in Case 2. The results show the remarkable energy-saving effects of Case 4. 4.1.3. Additional coal consumption rate due to the increment of flue gas pressure drop In practical engineering operations, flue gas pressure drop is increased by installing a heat exchanger in the tail flue, which leads to fan work increment. As a consequence, the coal consumption rate of a power unit is also increased. Thus, additional coal consumption rates and consequent effects caused by the increment of flue gas pressure drop in each scheme should be taken into consideration. Combining Eqs. (18)–(20) and considering the fan work caused s by flue gas pressure drop, the increased coal consumption rate bf is calculated as follows: s
s
bf ¼ b
Nf ðg=kW hÞ Pnet þ Nf
ð21Þ
Therefore, the variation curve of increased coal consumption caused by fan work is deduced (Fig. 12). 4.1.4. Net coal conservation rate After comprehensive consideration of the gross coal conservation rate due to waste heat utilization (Fig. 11) and the additional coal consumption rate caused by the increment of flue gas pressure drop (Fig. 12), the variation rule of net coal conservation rate is obtained (Fig. 13). From Figs. 11–13, we can find that: (1) with the drop of exhaust gas temperature Tex, gross coal conservation rate, the additional coal consumption rate and net coal conservation rate of all cases
Gross coal conservation rate, g/KWh
914
3.5 case 4
3.0 2.5 2.0
case 1
1.5
case 3 case 2
1.0 0.5 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 11. Variation curves of gross coal conservation rate.
915
0.7
case 4 0.6 0.5 0.4 0.3
case 1 0.2
case 3 case 2
0.1 110
105
100
95
90
Annual coal-saving revenue, Million USD
Additional coal consumption rate, g/kWh
G. Xu et al. / Applied Energy 112 (2013) 907–917
1.8
case 4
1.6 1.4 1.2 1.0
case 1
0.8
case 3
0.6
case 2
0.4 0.2 110
Exhaust gas temperature, ºC
105
100
95
90
85
Exhaust gas temperature, ºC Fig. 12. Variation curves of additional coal consumption rate. Fig. 14. Variation curves of annual coal-saving revenue.
case 4
2.5
2.0 case 1
1.5
case 3
1.0 case 2
shown in Fig. 14. The basic economic assumptions employed here include: (1) The assumed coal price, 120 USD per ton standard coal (750 CNY per ton standard coal), was the average cost to China electric generators in 2012. (2) The exchange rate is set as 6.25 CNY/USD. (3) The annual utilization hours of the unit are set as 5000 h. From Fig. 14, we can see that the annual coal-saving revenue and the net coal conservation rate have similar variation curves. The annual amount and revenue of coal saving in Case 4 can reach nearly 13,000 tons and 1.56 million USD per year with minimal exhaust gas temperatures (90 °C).
0.5
110
105
100
95
90
Exhaust gas temperature, ºC Fig. 13. Variation curves of net coal conservation rate.
will increase; (2) the gross coal conservation rate, additional coal consumption and net coal-saving consumption of the high-stage steam substitute scheme. (Case 4) is much larger than other schemes; (3) for gross coal conservation rate curve, all of the four cases increases linearly with the decline of Tex. Case 4 is the highest case. As shown in Fig. 11, when Tex equals to 90 °C, the gross coal conservation rate of Case 4 can reach 3.27 g/kW h, which is nearly 1.79 times of that in Case 1 and 3.15 times of that in Case 2. As shown in Fig. 11, (4) for additional coal consumption rate curve, all of the four cases increases rapidly with the decline of Tex. Among them, Case 4 is also the highest one and exponential growing with the increment of Tex. As shown in Fig. 12, when Tex equals to 90 °C, the additional coal conservation rate of Case 4 can reach 0.657 g/ kW h, which is nearly 2.63 times of that in Case 1 and 4.63 times of that in Case 2, and (5) for net coal conservation rate curve, all of the four cases increases with the decline of Tex. As shown in Fig. 13, Case 4 still ranks first among the cases. But its curve is obviously more flat than other cases. When Tex equals to 90 °C, the net coal conservation rate of Case 4 can reach 0.657 g/kW h, which is nearly 1.65 times of that in Case 1 and 2.96 times of that in Case 2.
4.2.2. Analysis of annual capital cost caused by waste heat recovery According to data of similar domestic retrofitting projects of utility boilers in China, the capital investment of a unit heat transfer area is approximately 120 USD/m2 [12]. Therefore, according to the heat transfer areas of various cases (Fig. 8), we calculated the total investment cost (Fig. 15). To evaluate the costs and benefits of the engineering project in different periods, we convert capital cost using equivalent conversion. According to the model of equivalent payment calculation, we introduce the capital recovery factor (CRF) concept [2,19,20]:
CRF ¼ ½i ð1 þ iÞn =½ð1 þ iÞn 1
ð22Þ
7
case 4 Total investment, Million USD
Net coal conservation rate, g/kWh
3.0
6 5 4 3
case 1 2
case 3 case 2
1
4.2. Techno-economic analysis of heat recovery system 110
4.2.1. Annual coal-saving revenue due to waste heat recovery Based on the above analysis of net coal conservation rate, the annual coal-saving revenue of various cases can be obtained, as
105
100
95
90
Exhaust gas temperature, ºC Fig. 15. Variation curves of total investment of heat recovery system.
G. Xu et al. / Applied Energy 112 (2013) 907–917
Cost Inv ¼ CRF Inv
ð23Þ
where ‘Inv’ refers to the total investment capital. Fig. 16 illustrates the variation curve of annual capital cost in the four cases. The annual capital cost has a similar variation curve with the total investment (Fig. 15). With the drop of exhaust gas temperature Tex, both total investment and annual capital cost increase in all cases. However, total investment and annual capital cost in Case 4 still ranks first and increase fastest with decline of exhaust gas temperature (Figs. 15 and 16). The reason mainly lies in that the heat exchanger of Case 4 has the lowest temperature differential and the largest heat transfer area (Fig. 8). As shown in Fig. 16, when exhaust gas temperature Tex reaches 90 °C, the annual capital cost in Case 4 reaches the maximum, which is nearly 2.64 times of that in Case 1 and 4.22 times of that in Case 2. 4.2.3. Analysis of annual operation and maintenance (O&M) cost due to waste heat recovery To simplify O&M cost calculation, we assumed that annual O&M cost is proportional to the total investment cost (Inv) with proportionality coefficient aOM, i.e., CostOM = aOM Inv. According to related literature [19], aOM can be set as 4%, that is, the O&M costs are fixed at 4% of the total investment cost per year. Fig. 17 illustrates the variation curves of annual O&M in the four cases. It can be seen that, the variation curves of CostOM in Fig. 17 correspond to that of total capital requirement in Fig. 15. 4.2.4. Comprehensive economic analysis and optimization Based on the aforementioned analysis, we obtain the net annual revenue (NAR) as follows:
NAR ¼ Incomecoal Cap Cost OM
ð24Þ
Here CostOM represents the O&M cost of the waste recovery system. Fig. 18 illustrates the variation curves of NAR for the heat recovery system. As seen from this figure, Case 4 still ranks first among the cases. However, the NAR of Case 4 will increase to an optimal point before decreasing as exhaust gas temperature decreases. This is owing to the fact that, with the drop of exhaust gas temperatures, coal-saving rate and revenue increase greatly (Figs. 13 and 14), but on the other hand, heat exchange areas and capital cost
Annual Capital Cost, Million USD
0.7
case 4 0.6 0.5 0.4 0.3
case 1 0.2
case 3 case 2
0.1 0.0 110
105
100
95
Exhaust gas temperature, ºC Fig. 16. Variation curves of annual capital cost.
90
0.30
Annual O&M Cost, Million USD
where i refers to the fraction interest rate per year, and n refers to the number of years that the capital has been borrowed over a fixed rate of interest. In this paper, i = 0.08 and n = 20. Therefore, the annual capital cost (CostInv) can be calculated as follows:
case 4 0.25
0.20
0.15
case 1
0.10
case 3 case 2
0.05
0.00
110
105
100
95
90
Exhaust gas temperature, ºC Fig. 17. Variation curves of annual O&M cost.
1.0
Net annual revenue, Million USD
916
optimal
0.9 0.8 0.7
case 4
0.6
case 1
0.5
case 3 0.4
case 2
0.3 0.2 0.1 110
105
100
95
90
Exhaust gas temperature, ºC Fig. 18. Variation curves of net annual revenue.
simultaneously increase on a large scale (Figs. 9 and 16), which counteracts part of coal-saving benefits. Finally, when temperature decrease reaches a certain value, the increasing capital cost will exceed the coal-saving benefits. As a consequence, the comprehensive performance curve will increase rather than decrease with variation of exhaust gas temperature. The optimal exhaust gas temperature is 98 °C, which enables the NAR to reach 0.889 million USD. Therefore, there is an interesting result. That is, from the techno-economic perspective, the performance of the high-stage steam substitute scheme of flue gas heat recovery system will not always increase with the increment of the recovered heat (the drop of exhaust gas temperature). Instead, there is an optimum point. Fig. 19 illustrates the variation rule of the optimal exhaust gas temperature, annual capital cost, annual O&M cost, annual coalsaving revenue and net annual revenue (NAR) in the best scheme (Case 4) when the coal price varies. As seen in Fig. 19, with the increase of coal price, the optimal exhaust flue gas temperature will gradually decrease and annual coal-saving revenue will increase dramatically. Meanwhile, the annual capital and O&M cost increases gently while the net annual revenue increases greatly. This is due to that with the increase of coal price, the lower exhaust gas temperature is suitable for the waste heat recovery system, as a consequence of which, more waste heat recovery and
104
Annual coal-saving revenue Optimal exhaust gas temperature
2.0
1.6
102 Net annual revenue
1.2 100 0.8 98 Annual capital cost
96
0.4
Annual O&M cost
Economic parameters, Million USD
Optimal exhaust gas temperature, ºC
G. Xu et al. / Applied Energy 112 (2013) 907–917
0.0 80
100
120
140
160
Standard coal price, USD/ton Fig. 19. Sensitivity analysis with standard coal price.
more energy-saving effects can be achieved. Meanwhile, the further decreased exhaust temperature of flue gas will lead to more heat surface, resulting in more annual capital and O&M cost. However, the increment of investment and O&M cost is obviously much less than that of annual coal-saving revenue, therefore, leads to the obvious increment of the net annual revenues. Thus, when the coal price is comparatively high, it is proper to adopt the design with more heat surfaces to recover more waste heat, as a consequence of which, the economic benefits caused by waste heat recover will also increase dramatically. 5. Conclusion We conducted a techno-economic analysis and optimization design of a heat recovery system using boiler exhaust gas. This study is based on the principles of thermodynamic engineering, heat transfer, and hydrodynamics together with data from an existing typical 1000 MW power generation unit in China. From the results of this study, important conclusions are obtained and a few interesting integration measures are proposed. (1) The effect of flue gas exhaust energy is affected not only by exhaust gas temperature in the boiler side but also by the integrated solution in the steam/water side. The grade of substitute extraction steam has a determinative influence on the energy-saving effect. (2) High-stage steam substitute scheme of flue gas heat recovery in power plant can obtain higher energy-saving effects than that of low-stage steam substitute scheme. (3) The energy-saving benefit, especially with high-stage steam substitute scheme, is weakened by large heat exchange areas and flue gas pressure drop caused by lower temperature differentials. (4) Techno-economic analysis reveals that the economic performance of the high-stage steam substitute scheme will not always increase with the increment of the recovered heat and have an optimum point. When the coal price is comparatively high, it is proper to adopt the design with more heat surfaces to recover more waste heat, as a consequence of which, the economic benefits caused by waste heat recover will also increase dramatically.
917
(5) Based on the data from an existing 1000 MW typical power generation unit in China, the annual amount and revenue of coal saving in an optimal high-stage steam substitute scheme (Case 4) can reach nearly 13,000 tons and 1.56 million USD per year with minimal exhaust gas temperatures (90 °C). However, the net annual revenue of the optimal high-stage steam substitute scheme (Case 4) can reach highest value (0.889 million USD) when the exhaust gas temperature is 98 °C.
Acknowledgements This paper is supported by the National Major Fundamental Research Program of China (2011CB710706), the National Nature Science Fund of China (U1261210, 51006034) and the 111 Project (B12034).
References [1] Liu ZY. Annual report on the development of China’s electric power industry. Beijing: Electricity Council; 2011 [in Chinese]. [2] Westerlund Lars, Hermansson Roger, Fagerström Jonathan. Flue gas purification and heat recovery: a biomass fired boiler supplied with an open absorption system. Appl Energy 2012;96:444–50. [3] Zhang JH, Zhou YL, Li Y, et al. Generalized predictive control applied in waste heat recovery power plants. Appl Energy 2013;102:320–6. [4] Blarke Morten B. Towards an intermittency-friendly energy system: comparing electric boilers and heat pumps in distributed cogeneration. Appl Energy 2012;91:349–65. [5] Chen Q, Finney Karen, Li HN, et al. Condensing boiler applications in the process industry. Appl Energy 2012;89:30–6. [6] Pandiyarajan V, Chinna Pandian M, Malan E, et al. Experimental investigation on heat recovery from diesel engine exhaust using finned shell and tube heat exchanger and thermal storage system. Appl Energy 2011;88:77–87. [7] Espatolero Sergio, Cortés Cristóbal, Romeo Luis M. Optimization of boiler coldend and integration with the steam cycle in supercritical units. Appl Energy 2010;87:1651–60. [8] Jeong K, Kessen MJ, Bilirgen H. Analytical modeling of water condensation in condensing heat exchanger. Int J Heat Mass Transfer 2010;53:2361–8. [9] Jannelli E, Minutillo M. Simulation of the flue gas cleaning system of an RDF incineration power plant. Waste Manage 2007;27:684–90. [10] Wang CJ, He BS, Sun SY, et al. Application of a low pressure economizer for waste heat recovery from the exhaust flue gas in a 600 MW power plant. Energy 2012 [online 20 February 2012]. [11] Wang DX, Bao A, Kunc W, et al. Coal power plant flue gas waste heat and water recovery. Appl Energy 2012;91:341–8. [12] Zhao ZJ, Feng WZ, Zhang L, et al. Theoretical analysis and engineering practice of heat recovery form exhaust gas of power boilers. J Power Eng 2009;29:994–7 [in Chinese]. [13] Lin WC. Energy-saving theory of power plant system. Xi’an: Xi’an Jiaotong University Press; 1994 [in Chinese]. [14] Bahadori A. Estimation of combustion flue gas acid dew point during heat recovery and efficiency gain. Appl Therm Eng 2011;31:1457–62. [15] Manassaldi JI, Mussati SF, Scenna NJ. Optimal synthesis and design of heat recovery steam generation (HRSG) via mathematical programming. Energy 2011;36:475–85. [16] Yang SM, Tao WQ. Heat transfer. Beijing: Higher Education Press; 2006 [in Chinese]. [17] Kiyoshi K, Kenichi O, Takaharu K. Heat transfer and pressure drop characteristics of finned tube banks in forced convection (comparison of the pressure drop characteristics between spiral fin and serrated fin). Heat Transfer – Asian Res 2004;33(7):431–44. [18] Weierman C. Correlations ease the selection of fin tubes. Oil Gas J 1976;74(6):94–100. [19] Kreutz T, Williams R, Consonni S, Chiesa P. Co-production of hydrogen, electricity, and CO2 from coal with commercially ready technology. Part B: economic analysis. Hydrogen Energy 2005;30(7):769–84. [20] Xu G, Jin HG, Yang YP, et al. A comprehensive techno-economic analysis method for power generation systems with CO2 capture. Int J Energy Res 2010;34(4):321–32.