All about steam

Mr. C.S Joshi
Director, FARE Labs Pvt Ltd. Gurugram, Haryana.India

Email: csjoshi@farelabs.com

• Conversion of fuel into steam:

Calorific Value is a crucial factor in determining the amount of steam a boiler can generate, as it directly relates to the energy input available for converting water into steam. A higher calorific value means more heat is produced per unit of fuel, potentially leading to more steam generation. 

Factors Affecting Steam Generation:

• Calorific Value:

This is the primary factor. Fuels with higher calorific values (like certain types of coal or oil) will produce more heat per unit mass, requiring less fuel for the same steam output. 

• Boiler Efficiency:

This refers to how effectively the boiler converts the heat from the fuel into steam. A more efficient boiler will produce more steam from the same amount of fuel. 

• Fuel Type:

Different fuels have different calorific values. For example, fuel oils generally have more consistent calorific values than coal, which can vary widely depending on its type and composition. 

• Steam Requirements:

The desired temperature, pressure, and flow rate of the steam will affect the amount of fuel needed. 

• Boiler Design and Operation:

Factors like the type of boiler, the way it’s operated, and the heat transfer efficiency all play a role in determining how much steam is produced from a given amount of fuel. 

Fuel-to-steam efficiency = (Energy output by steam / Energy input by fuel) * 100%

• Energy input by fuel: Fuel consumption * Calorific value of fuel
• Energy output by steam: Steam flow * (Steam enthalpy – Feed water enthalpy)

Q = 540 Kcal is required to convert 1 kg of water to steam.

4.         Steam Table:

Steam and temperature possess a reliable relationship whereby if the pressure of the steam is known, its temperature can be predicted (and vice versa). Below is a graph and table of this relationship.

Saturated Steam – Properties vs. Pressure – Bar
Absolute Pressure	Boiling Point	Specific Volume (steam)	Density (steam)	Specific Enthalpy of Liquid Water (sensible heat)		Specific Enthalpy of Steam (total heat)		Latent heat of Vaporization		Specific Heat
(bar)	(oC)	(m3/kg)	(kg/m3)	(kJ/kg)	(kcal/kg)	(kJ/kg)	(kcal/kg)	(kJ/kg)	(kcal/kg)	(kJ/kgK)
0.02	17.51	67.006	0.015	73.45	17.54	2533.64	605.15	2460.19	587.61	1.8644
0.03	24.10	45.667	0.022	101.00	24.12	2545.64	608.02	2444.65	583.89	1.8694
0.04	28.98	34.802	0.029	121.41	29.00	2554.51	610.13	2433.10	581.14	1.8736
0.05	32.90	28.194	0.035	137.77	32.91	2561.59	611.83	2423.82	578.92	1.8774
0.06	36.18	23.741	0.042	151.50	36.19	2567.51	613.24	2416.01	577.05	1.8808
0.07	39.02	20.531	0.049	163.38	39.02	2572.62	614.46	2409.24	575.44	1.8840
0.08	41.53	18.105	0.055	173.87	41.53	2577.11	615.53	2403.25	574.01	1.8871
0.09	43.79	16.204	0.062	183.28	43.78	2581.14	616.49	2397.85	572.72	1.8899
0.1	45.83	14.675	0.068	191.84	45.82	2584.78	617.36	2392.94	571.54	1.8927
0.2	60.09	7.650	0.131	251.46	60.06	2609.86	623.35	2358.40	563.30	1.9156
0.3	69.13	5.229	0.191	289.31	69.10	2625.43	627.07	2336.13	557.97	1.9343
0.4	75.89	3.993	0.250	317.65	75.87	2636.88	629.81	2319.23	553.94	1.9506
0.5	81.35	3.240	0.309	340.57	81.34	2645.99	631.98	2305.42	550.64	1.9654
0.6	85.95	2.732	0.366	359.93	85.97	2653.57	633.79	2293.64	547.83	1.9790
0.7	89.96	2.365	0.423	376.77	89.99	2660.07	635.35	2283.30	545.36	1.9919
0.8	93.51	2.087	0.479	391.73	93.56	2665.77	636.71	2274.05	543.15	2.0040
0.9	96.71	1.869	0.535	405.21	96.78	2670.85	637.92	2265.65	541.14	2.0156
1 1)	99.63	1.694	0.590	417.51	99.72	2675.43	639.02	2257.92	539.30	2.0267
1.1	102.32	1.549	0.645	428.84	102.43	2679.61	640.01	2250.76	537.59	2.0373
1.2	104.81	1.428	0.700	439.36	104.94	2683.44	640.93	2244.08	535.99	2.0476
1.3	107.13	1.325	0.755	449.19	107.29	2686.98	641.77	2237.79	534.49	2.0576
1.4	109.32	1.236	0.809	458.42	109.49	2690.28	642.56	2231.86	533.07	2.0673
1.5	111.37	1.159	0.863	467.13	111.57	2693.36	643.30	2226.23	531.73	2.0768
Saturated Steam – Properties vs. Pressure – Bar
Absolute Pressure	Boiling Point	Specific Volume (steam)	Density (steam)	Specific Enthalpy of Liquid Water (sensible heat)		Specific Enthalpy of Steam (total heat)		Latent heat of Vaporization		Specific Heat
(bar)	(oC)	(m3/kg)	(kg/m3)	(kJ/kg)	(kcal/kg)	(kJ/kg)	(kcal/kg)	(kJ/kg)	(kcal/kg)	(kJ/kgK)
1.6	113.32	1.091	0.916	475.38	113.54	2696.25	643.99	2220.87	530.45	2.0860
1.7	115.17	1.031	0.970	483.22	115.42	2698.97	644.64	2215.75	529.22	2.0950
1.8	116.93	0.977	1.023	490.70	117.20	2701.54	645.25	2210.84	528.05	2.1037
1.9	118.62	0.929	1.076	497.85	118.91	2703.98	645.83	2206.13	526.92	2.1124
2	120.23	0.885	1.129	504.71	120.55	2706.29	646.39	2201.59	525.84	2.1208
2.2	123.27	0.810	1.235	517.63	123.63	2710.60	647.42	2192.98	523.78	2.1372
2.4	126.09	0.746	1.340	529.64	126.50	2714.55	648.36	2184.91	521.86	2.1531
2.6	128.73	0.693	1.444	540.88	129.19	2718.17	649.22	2177.30	520.04	2.1685
2.8	131.20	0.646	1.548	551.45	131.71	2721.54	650.03	2170.08	518.32	2.1835
3	133.54	0.606	1.651	561.44	134.10	2724.66	650.77	2163.22	516.68	2.1981
3.5	138.87	0.524	1.908	584.28	139.55	2731.63	652.44	2147.35	512.89	2.2331
4	143.63	0.462	2.163	604.68	144.43	2737.63	653.87	2132.95	509.45	2.2664
4.5	147.92	0.414	2.417	623.17	148.84	2742.88	655.13	2119.71	506.29	2.2983
5	151.85	0.375	2.669	640.12	152.89	2747.54	656.24	2107.42	503.35	2.3289
5.5	155.47	0.342	2.920	655.81	156.64	2751.70	657.23	2095.90	500.60	2.3585
6	158.84	0.315	3.170	670.43	160.13	2755.46	658.13	2085.03	498.00	2.3873
6.5	161.99	0.292	3.419	684.14	163.40	2758.87	658.94	2074.73	495.54	2.4152
7	164.96	0.273	3.667	697.07	166.49	2761.98	659.69	2064.92	493.20	2.4424
7.5	167.76	0.255	3.915	709.30	169.41	2764.84	660.37	2055.53	490.96	2.4690
8	170.42	0.240	4.162	720.94	172.19	2767.46	661.00	2046.53	488.80	2.4951
8.5	172.94	0.227	4.409	732.03	174.84	2769.89	661.58	2037.86	486.73	2.5206
9	175.36	0.215	4.655	742.64	177.38	2772.13	662.11	2029.49	484.74	2.5456
9.5	177.67	0.204	4.901	752.82	179.81	2774.22	662.61	2021.40	482.80	2.5702
10	179.88	0.194	5.147	762.60	182.14	2776.16	663.07	2013.56	480.93	2.5944
11	184.06	0.177	5.638	781.11	186.57	2779.66	663.91	1998.55	477.35	2.6418
12	187.96	0.163	6.127	798.42	190.70	2782.73	664.64	1984.31	473.94	2.6878
Saturated Steam – Properties vs. Pressure – Bar
Absolute Pressure	Boiling Point	Specific Volume (steam)	Density (steam)	Specific Enthalpy of Liquid Water (sensible heat)		Specific Enthalpy of Steam (total heat)		Latent heat of Vaporization		Specific Heat
(bar)	(oC)	(m3/kg)	(kg/m3)	(kJ/kg)	(kcal/kg)	(kJ/kg)	(kcal/kg)	(kJ/kg)	(kcal/kg)	(kJ/kgK)
13	191.60	0.151	6.617	814.68	194.58	2785.42	665.29	1970.73	470.70	2.7327
14	195.04	0.141	7.106	830.05	198.26	2787.79	665.85	1957.73	467.60	2.7767
15	198.28	0.132	7.596	844.64	201.74	2789.88	666.35	1945.24	464.61	2.8197
16	201.37	0.124	8.085	858.54	205.06	2791.73	666.79	1933.19	461.74	2.8620
17	204.30	0.117	8.575	871.82	208.23	2793.37	667.18	1921.55	458.95	2.9036
18	207.11	0.110	9.065	884.55	211.27	2794.81	667.53	1910.27	456.26	2.9445
19	209.79	0.105	9.556	896.78	214.19	2796.09	667.83	1899.31	453.64	2.9849
20	212.37	0.100	10.047	908.56	217.01	2797.21	668.10	1888.65	451.10	3.0248
21	214.85	0.095	10.539	919.93	219.72	2798.18	668.33	1878.25	448.61	3.0643
22	217.24	0.091	11.032	930.92	222.35	2799.03	668.54	1868.11	446.19	3.1034
23	219.55	0.087	11.525	941.57	224.89	2799.77	668.71	1858.20	443.82	3.1421
24	221.78	0.083	12.020	951.90	227.36	2800.39	668.86	1848.49	441.50	3.1805
25	223.94	0.080	12.515	961.93	229.75	2800.91	668.99	1838.98	439.23	3.2187
26	226.03	0.077	13.012	971.69	232.08	2801.35	669.09	1829.66	437.01	3.2567
27	228.06	0.074	13.509	981.19	234.35	2801.69	669.17	1820.50	434.82	3.2944
28	230.04	0.071	14.008	990.46	236.57	2801.96	669.24	1811.50	432.67	3.3320
29	231.96	0.069	14.508	999.50	238.73	2802.15	669.28	1802.65	430.56	3.3695
30	233.84	0.067	15.009	1008.33	240.84	2802.27	669.31	1793.94	428.48	3.4069

1) 1 bar abs = 0 bar gauge = 0.98692 atm = 100 kPa abs = Standard Atmospheric Pressure (IUPAC system)

•             Vacuum steam is the general term used for saturated steam at temperatures below 100°C.

•             One calorie (kcal) equals 4.184 kJ or 4,184 joules (J).

Example: Boiling Water at 100 °C, 0 bar (100 kPa) Atmospheric Pressure

At atmospheric pressure (0 bar g, absolute 1 bar) water boils at 100 °C and 417.51 kJ of energy is required to heat 1 kg of water from 0 °C to evaporating temperature 100 °C.

Therefore the specific enthalpy of water at 0 bar g (absolute 1 bar) and 100 oC is 417.51 kJ/kg.

Another 2257.92 kJ of energy is required to evaporate 1 kg of water at 100 oC into 1 kg of steam at 100 oC. Therefore at 0 bar g (absolute 1 bar) the specific enthalpy of vaporation is 2257.19 kJ/kg.

The total specific enthalpy for steam at 0 bar gauge is:

hs = (417.51 kJ/kg) + (2257.92 kJ/kg) = 2675.43 kJ/kg = 640.1 kcal/kg 

Example: Boiling Water at 170 °C and 7 bar (700 kPa) Atmospheric Pressure

At 7 bar g (absolute 8 bar) the saturation temperature of water is 170.42 °C. More heat energy is required to raise its temperature to saturation point at 7 bar g than needed when the water is at atmospheric pressure. According to the table 720.94 kJ is required to raise 1 kg of water from 0 °C to saturation temperature 170 °C.

The heat energy (enthalpy of evaporation) needed at 7 bar g to vaporize the water to steam is actually less than required at atmospheric pressure. The specific enthalpy of vaporization decreases with steam pressure. The evaporation heat is 2046.53 kJ/kg at 7 bar g.

 The specific volume of steam decreases with increased pressure – and the amount of heat energy distributed by the same volume increase. With higher pressure – more energy can be transferred in a steam distribution system.

Example: At atmospheric pressure (10 bar g, absolute 11 bar) water boils at 179.88  °C and 762.6 kJ of energy is required to heat 1 kg of water from 0 °C to evaporating temperature 179.88   °C

Therefore the specific enthalpy of water at 10 bar g (absolute 11 bar) and 179.88 oC is 762.6 kJ/kg.

Another 2013.56 kJ of energy is required to evaporate 1 kg of water at 179.88 °C into 1 kg of steam at 179.88 °C. Therefore at 10 bar g (absolute 11 bar) the specific enthalpy of vaporization is 2013.56 kJ/kg.

The total specific enthalpy for steam at 10 bar gauge is:

hs = (762.6 kJ/kg) + (2013.56 kJ/kg)  = 2776.16 kJ/kg = 663.07 kcal/kg 

5. Selection of steam pipelines in process plant:

Sizing of Steam Pipes
Capacity (kg/h)
Gauge Pressure (bar)	Steam Velocity (m/s)	Nominal Pipe Size (mm)
		15	20	25	32	40	50	65	80	100	125	150
		Inside Diameter (mm)
		15.8	20.9	26.6	35.0	40.9	52.5	62.7	77.9	102.3	128.2	154.1
0.4	15	9	15	24	42	57	95	135	208	359	564	815
	25	14	25	41	70	96	158	225	347	598	940	1358
	40	23	40	65	112	153	252	360	555	958	1504	2173
0.5	15	9	16	26	45	61	101	144	222	383	602	869
	25	15	27	43	75	102	168	240	370	638	1003	1449
	40	24	43	69	120	163	269	384	592	1021	1604	2318
Sizing of Steam Pipes
Capacity (kg/h)
Gauge Pressure (bar)	Steam Velocity (m/s)	Nominal Pipe Size (mm)
		15	20	25	32	40	50	65	80	100	125	150
		Inside Diameter (mm)
		15.8	20.9	26.6	35.0	40.9	52.5	62.7	77.9	102.3	128.2	154.1
0.7	15	10	18	29	51	69	113	162	250	431	676	977
	25	17	30	49	84	115	189	270	416	718	1127	1628
	40	27	48	78	135	184	302	431	666	1148	1803	2605
0.9	15	11	20	32	56	76	126	179	277	478	750	1084
	25	19	33	54	93	127	210	299	462	796	1250	1806
	40	30	53	86	149	204	335	478	738	1274	2000	2890
1	15	12	21	34	59	80	132	188	291	501	787	1137
	25	20	35	57	98	133	220	314	484	835	1312	1895
	40	32	56	91	157	214	352	502	775	1336	2099	3032
2	15	17	31	50	86	117	193	275	425	733	1151	1663
	25	29	51	83	143	195	322	459	708	1221	1918	2771
	40	47	82	133	229	312	515	734	1133	1954	3069	4434
3	15	23	40	65	113	153	253	361	557	960	1508	2178
	25	38	67	109	188	256	421	601	928	1600	2513	3631
	40	61	107	174	300	409	674	962	1485	2560	4021	5809
4	15	28	50	80	139	189	312	445	687	1185	1860	2688
	25	47	83	134	232	316	520	742	1145	1974	3101	4480
	40	75	132	214	371	505	832	1187	1832	3159	4961	7168
5	15	34	59	95	165	225	371	529	816	1407	2210	3193
	25	56	98	159	275	375	618	881	1360	2345	3683	5321
	40	90	157	254	440	600	988	1409	2176	3752	5892	8514
6	15	39	68	110	191	260	429	611	944	1628	2556	3693
	25	65	114	184	318	434	714	1019	1573	2713	4260	6155
	40	104	182	294	509	694	1143	1630	2517	4340	6816	9848
Sizing of Steam Pipes
Capacity (kg/h)
Gauge Pressure (bar)	Steam Velocity (m/s)	Nominal Pipe Size (mm)
		15	20	25	32	40	50	65	80	100	125	150
		Inside Diameter (mm)
		15.8	20.9	26.6	35.0	40.9	52.5	62.7	77.9	102.3	128.2	154.1
7	15	44	77	125	217	295	487	694	1071	1847	2901	4192
	25	73	129	209	361	492	811	1157	1785	3079	4835	6986
	40	118	206	334	578	787	1297	1851	2856	4926	7736	11178
8	15	49	86	140	242	330	544	776	1198	2066	3245	4688
	25	82	144	234	404	550	907	1294	1997	3444	5408	7814
	40	131	231	374	646	881	1451	2070	3195	5510	8653	12502
10	15	60	105	170	294	400	659	940	1451	2502	3930	5678
	25	99	175	283	489	667	1098	1567	2418	4171	6550	9464
	40	159	279	453	783	1067n	1758	2507	3869	6673	10480	15142
14	15	80	141	229	396	539	888	1266	1955	3371	5295	7650
	25	134	235	381	659	898	1480	2111	3258	5619	8825	12750
	40	214	376	610	1055	1437	2368	3377	5213	8991	14119	20401

The values above are calculated for steel pipes medium thickness schedule 40. Using other standards with different inside diameters, the values above should be compensated.

Example – Capacity of 50 mm Steam Pipe

The capacity of a 50 mm steam pipe with 10 bar pressure and steam velocity 25 m/s is 1098 kg/h.

Steam pipelines should be decided as per above table for best economy of steam consumption.

6.         Coil size for steam heat transfer:

1. Heating with coils and Jackets

Indirect heating of fluids is covered in this tutorial including layouts, control and drainage of coils and jackets, and heat transfer calculations. Vessels can be heated in a few different ways. We will deal with indirect heating. In these systems, the heat is transferred across a heat transfer surface. Options include:

• Submerged steam coils – A widely used form of heat transfer involves the installation inside a tank of a steam coil immersed in a process fluid.
• Steam jackets-Steam circulates in the annular space space between a jacket and the vessel walls, and heat is transferred through the wall of the vessel.

SUBMERGED STEAM COILS

The use of tank coils is particularly common in marine applications where cargoes of crude oil, edible oils, tallow and molasses are heated in deep tanks. Many of these liquids are difficult to handle at ambient temperatures due to their viscosity. Steam heated coils are used to raise the temperature of these liquids, lowering their viscosity so that they become easier to pump.

Tank coils are also extensively used in electroplating and metal treatment. Electroplating involves passing articles through several process tanks so that metallic coatings can be deposited on to their surfaces. One of the first stages in this process is known as pickling, where materials such as steel and copper are treated by dipping them in tanks of acid or caustic solution to remove any scale or oxide (e.g. rust) which may have formed.

Steam coil sizing

Having determined the energy required, and with knowledge of the steam pressure/temperature in the coil, the heat transfer surface may be determined using Equation 2.5.3:

The heat transfer area calculated is equivalent to the surface area of the coil, and will enable an appropriate size and layout to be specified.

Determining the ‘U’ value

To calculate the heat transfer area, a value for the overall heat transfer coefficient, U, must be chosen. This will vary considerably with the thermal and transport properties of both fluids and a range of other conditions.

On the product side of the coil a thermal boundary layer will exist in which there is a temperature gradient between the surface and the bulk fluid. If this temperature difference is relatively large, then the natural convective currents will be significant, and the heat transfer coefficient will be high.

Assisted circulation (such as stirring) that will induce forced convection, will also result in higher coefficients. As convection is partially dependent on the bulk motion of the fluid, the viscosity (which varies with temperature) also has an important bearing on the thermal boundary layer.

Additional variations can also occur on the steam side of the coil, especially with long lengths of pipe. The coil inlet may have a high steam velocity and may be relatively free from water.
However, further along the length of the coil the steam velocity may be lower, and the coil may be running partially full of water. In very long coils, such as those sometimes found in seagoing tankers or in large bulk storage tanks, a significant pressure drop occurs along the length of the coil. To achieve the mean coil temperature, an average steam pressure of approximately 75% of the inlet pressure may be used. In extreme cases the average pressure used may be as low as 40% of the inlet pressure. Another variable is the coil material itself. The thermal conductivity of the coil material may vary considerably.

However, overall heat transfer is governed to a large extent by the heat-resistant films, and the thermal conductivity of the coil material is not as significant as their combined effect. Table 2.10.1 provides typical overall heat transfer coefficients for various conditions of submerged steam coil application. ‘U’ values for steam pressures between 2 bar g and 6 bar g should be found by interpolation of the data in the table.

Table 2.10.1 Heat emission rates for steam coils submerged in water

Customary overall heat transfer coefficients		U (W/m² °C)
Mean steam/water temperature difference around 30 °C		550 – 1 300
Mean steam/water temperature difference around 60 °C		1 000 – 1 700
Mean steam/water temperature difference around 110 °C		1 300 – 2 700
Recommended rates		U (W/m² °C)
Lower pressure coils	(<2 bar g) with natural circulation of water	550
Higher pressure coils	(>6 bar g) with natural circulation of water	1 100
Lower pressure coils	(<2 bar g) with assisted circulation of water	1 100
Higher pressure coils	(>6 bar g) with assisted circulation of water	1 700

The range of figures shown in Table 2.10.1 demonstrates the difficulty in providing definitive ‘U’ values. Customary figures at the higher end of the scale will apply to installations that are supplied with clean dry steam, small coils and good condensate drainage. The lower end is more applicable to poor quality steam, long coils and poor condensate drainage.

The recommended overall heat transfer coefficients will apply to typical conditions and installations.

These recommended rates are empirically derived, and will generally ensure that a generous safety margin applies to the coil sizing.

In the case of fluids other than water, the heat transfer coefficient will vary even more widely due to the way in which viscosity varies with temperature. However, the values shown in Table 2.10.2 will serve as a guide for some commonly encountered substances, while Table 2.10.3 gives typical surface areas of pipes per metre length.

Table 2.10.2 Heat emission rates for steam coils submerged in miscellaneous liquids

Medium pressure steam	(2 – 6 bar g) with natural liquid convection	U (W/m² °C)
Light oils		170
Heavy oils		80 – 110
Fats *		30 – 60
Medium pressure steam	(2 – 6 bar g) with forced liquid convection	U (W/m² °C)
Light oils	(200 sec Redwood at 38 °C)	550
Medium oils	(1 000 sec Redwood at 38 °C)	340
Heavy oils	(3 500 sec Redwood at 38 °C)	170
Molasses **	(10 000 sec Redwood at 38 °C)	85
Fats *	(50 000 sec Redwood at 38 °C)	55

* Certain materials such as tallow and margarine are solid at normal temperatures but have quite low viscosities in the molten state.
** Commercial molasses frequently contains water and the viscosity is much lower.

Table 2.10.3 Nominal surface areas of steel pipes per meter length

Nominal bore (mm)	15	20	25	32	40	50	65	80	100
Surface area (m²/m)	0.067	0.085	0.106	0.134	0.152	0.189	0.239	0.279	0.358

Example 2.​10.1

Example 2.9.1

For the tank shown in Figure 2.9.3, determine:

Part 1. The mean heat transfer rate required during start-up.

Part 2. The maximum heat transfer rate required during operation.

Fig 2.9.3 Overall heat transfer coefficients for oil tanks

The tank is unlagged and open topped and is situated on a concrete floor inside a factory.

It is 3 m long by 3 m wide by 2 m high.

Tank total surface area = 24 m² (excluding base).

Heat transfer coefficient from tank/air, U1 = 11 W/m² °C.

The tank is 2/3 full of a weak acid solution (cp = 3.9 kJ/kg °C) which has the same density as water (1 000 kg/m³)

The tank is fabricated from 15 mm mild steel plate. (Density = 7 850 kg/m³, cp = 0.5 kJ/kg °C)

The tank is used on alternate days, when the solution needs to be raised from the lowest considered ambient temperature of 8 °C to 60 °C in 2 hours, and remain at that temperature during the day.

When the tank is up to temperature, a 500 kg steel article is to be dipped every 20 minutes without the tank overflowing. (cp = 0.5 kJ/kg °C

Continuing from Example 2.9.1 determine:

• Part 1. The average steam mass flowrate during start-up. (Mean heat load = 367 kW)
• Part 2. The heat transfer area required
• Part 3. A recommended coil surface area.
• Part 4. The maximum steam mass flowrate with the recommended heat transfer area.
• Part 5. A recommendation for installation, including coil diameter and layout.

The following additional information has been provided:

• Steam pressure onto the control valve = 2.6 bar g (3.6 bar a).
• A stainless steel steam coil provides heat.
• Heat transfer coefficient from steam/coil/liquid, U = 650 W/m² °C

Part 1 Calculate the average steam mass flowrate during start-up

Steam pressure onto the control valve = 2.6 bar g (3.6 bar a)

Critical pressure drop (CPD) will occur across the control valve during start-up, therefore the minimum steam pressure in the heating coil should be taken as 58% of upstream absolute pressure. An explanation of this is given in Block 5.

Part 2 Calculate the heat transfer area required

Part 3 A recommendation for coil surface area

Because of the difficulties in providing accurate ‘U’ values, and to allow for future fouling of the heat exchange surface, it is usual to add 10% to the calculated heat transfer area.

Part 4 the maximum steam mass flowrate with the recommended heat transfer area

Maximum heat transfer (and hence steam demand) will occur when the temperature difference between the steam and the process fluid is at its maximum, and should take into consideration the extra pipe area allowed for fouling.

(a) Consider the maximum heating capacity of the coil Q̇(coil)

(b) Steam flowrate to deliver 519 kW

Part 5 A recommendation for installation, including coil diameter and layout

a) Determine coil diameter and length

It may be difficult to accommodate this length of large bore heating pipe to install in a 3 m × 3 m tank.

One solution would be to run a bank of parallel pipes between steam and condensate manifolds, set at different heights to encourage condensate to run to the lower (condensate) manifold. The drain line must fall from the bottom of the condensate manifold down to the steam trap (or pump-trap).

See Figure 2.10.1 for a suggested layout.

Note the steam supply is situated at one end of its manifold, whilst the trap set is at the other end. This will help steam to flow and push condensate through the coils.
In the application, the steam and condensate headers would each be 2.8 m long. As the condensate manifold is holding condensate, the heat from it will be small compared to the steam manifold and this can be ignored in the calculation.

The steam manifold should be 100 mm diameter as determined by the previous velocity calculation. This will provide a heating area of:

2.8 m x 0.358 m²/m = 1.0 m²

Consequently 7 m² – 1 m² = 6 m² of heat transfer area is still required and must be provided by the connecting pipes.

Arbitrarily selecting 32 mm pipe as a good compromise between robustness and workability:

It is necessary to confirm the steam velocity through the connecting tubes:

On the basis of proportionality of heat transfer area, the steam header will condense

Other Types of steam coil layouts

The design and layout of the steam coil will depend on the process fluid being heated. When the process fluid to be heated is a corrosive solution, it is normally recommended that the coil inlet and outlet connections are taken over the lip of the tank, as it is not normally advisable to drill through the corrosion resistant linings of the tank side. This will ensure that there are no weak points in the tank lining, where there is a risk of leakage of corrosive liquids. In these cases the coil itself may also be made of corrosion resistant material such as lead covered steel or copper, or alloys such as titanium.

However, where there is no danger of corrosion, lifts over the tank structure should be avoided, and the steam inlet and outlet connections may be taken through the tank side. The presence of any lift will result in waterlogging of a proportion of the coil length, and possibly water hammer, noise and leaking pipework.

Steam heating coils should generally have a gradual fall from the inlet to the outlet to ensure that condensate runs toward the outlet and does not collect in the bottom of the coil.

Where a lift is unavoidable, it should be designed to include a seal arrangement at the bottom of the lift and a small-bore dip pipe, as shown in Figure 2.10.2.

The seal arrangement allows a small amount of condensate to collect to act as a water seal, and prevents the occurrence of steam locking. Without this seal, steam can pass over any condensate collecting in the bottom of the pipe, and close the steam trap at the top of the riser.

The condensate level would then rise and form a temporary water seal, locking the steam between the bottom of the riser and the steam trap. The steam trap remains closed until the locked steam condenses, during which time the coil continues to waterlog.

When the locked steam condenses and the steam trap opens, a slug of water is discharged up the riser. As soon as the water seal is broken, steam will enter the rising pipe and close the trap, while the broken column of water falls back to lie at the bottom of the heating coil.

The small bore dip pipe will only allow a very small volume of steam to become locked in the riser. It enables the water column to be easily maintained without steam bubbling through it, ensuring there is a steady and continuous condensate flow to the outlet.

When the seal is ultimately broken, a smaller volume of water will return to the heating coil than with an unrestricted large bore riser, but as the water seal arrangement requires a smaller volume of condensate to form a water seal, it will immediately re-form.

If the process involves articles being dipped into the liquid, it may not be convenient to install the coil at the bottom of the tank – it may be damaged by the objects being immersed in the solution.
Also, during certain processes, heavy deposits will settle at the bottom of the tank and can quickly cover the heating surface, inhibiting heat transfer.

For these reasons side hung coils are often used in the electroplating industry. In such cases serpentine or plate-type coils are arranged down the side of a tank, as shown in Figure 2.10.3. These coils should also have a fall to the bottom with a water seal and a small bore dip-pipe.This arrangement has the advantage that it is often easier to install, and also easier to remove for periodic cleaning if required.

If articles are to be dipped into the tank, it may not be possible to use any sort of agitator to induce forced convection and prevent temperature gradients occurring throughout the tank. Whether bottom or side coils are used, it is essential that they are arranged with adequate coverage so that the heat is distributed evenly throughout the bulk of the liquid.

The diameter of the coil should provide sufficient length of coil for good distribution. A short length of coil with a large diameter may not provide adequate temperature distribution. However a very long continuous length of coil may experience a temperature gradient due to the pressure drop from end to end, resulting in uneven heating of the liquid.

• Control valve arrangement

The control valve set may be either one or two valves in parallel. A single control valve, large enough to cope with the maximum flowrate encountered at start-up, may be unable to control flow accurately at the minimum expected flowrate. This could cause erratic temperature control.
An alternative is to fit two temperature control valves in parallel:

• One valve (running valve) sized to control at the lower flowrate.
• A second valve (starting valve) to pass the difference between the capacity of the first valve, and the maximum flowrate.

The starting valve would have a set-point slightly lower than the running valve, so it would close first, leaving the running valve to control at low loads.

• Sizing the control valve

The control valve set (either one valve or two valves in parallel).

The coil has been sized on mean heat transfer values. However, it may be better to size the control valve to supply the maximum (start-up) load. With large coils in tanks, this will help to maintain a degree of steam pressure throughout the length of the coil when the steam is turned on, helping to push condensate through the coil to the steam trapping device. If the control valve were sized on mean values, steam pressure in the coil at start-up will tend to be lower and the coil may flood

Using one valve

Continuing with Example 2.10.1 the maximum steam load is 850 kg/h and the coil is designed to deliver this at a pressure of 1.1 bar g. A steam valve sizing chart would show that a Kv of about 20 is required to pass 850 kg/h of steam with a pressure of 2.6 bar g at the inlet of the control valve, and Critical Pressure Drop (CPD) across the valve. (Module 6.4 will show how the valve size can be determined by calculation).

A DN40 control valve with a larger Kvs of 25 would therefore need to be selected for the application.

If one valve is to be used, this valve must ensure the maximum heat load is catered for, while maintaining the required steam pressure in the coil to assist the drainage of condensate from it at start-up. However, for reasons previously explained, two valves may be better.

The running load is 52 kW and with the coil running at 1.1 bar g, the running steam load:

The steam valve sizing chart shows a Kv of 2 is required to pass 85 kg/h with 3.6 bar upstream, operating at critical pressure drop.

A DN15 KE type valve (Kvs = 4) and a DN25 piston actuated valve (Kvs = 18.6) operating together will cater for the start-up load. When approaching the control temperature, the larger valve would be set to shut down, allowing the smaller valve to give good control.

The condensate removal device

The selection and sizing of the condensate removal device will be very much influenced by the condensate backpressure. For the purpose of this example, it is assumed the backpressure is atmospheric pressure. The device should be sized so it is able to satisfy both of the following
conditions:

1. Pass 850 kg/h of condensate with 1.1 bar g in the coil, i.e. the full-load condition.
2. Pass the condensate load when steam pressure in the coil equals the condensate backpressure, i.e. the stall load condition.

If the steam trap is only sized on the first condition, it is possible that it may not pass the stall load (the condition where the product approaches its required temperature and the control valve modulates to reduce steam pressure). The stall load may be considerable. With respect to non-flow type applications such as tanks, this may not be too serious from a thermal viewpoint because the contents of the tank will almost be at the required temperature, and have a huge reservoir of heat.

Any reduction in heat transfer at this part of the heating process may therefore have little immediate effect on the tank contents.

However, condensate will back up into the coil and waterhammer will occur, along with its associated symptoms and mechanical stresses. Tank coils in large circular tanks tend to be of robust construction, and are often able to withstand such stresses. Problems can however occur in rectangular tanks (which tend to be smaller), where vibration in the coil will have more of an effect on the tank structure. Here, the energy dissipated by the waterhammer causes vibration, which can be detrimental to the life of the coil, the tank, and the steam trap, as well as creating unpleasant noise.

With respect to flow-type applications such as plate heat exchangers, a failure to consider the stall condition will usually have serious implications. This is mainly due to the small volume in the heat exchanger. For heat exchangers, any unwanted reduction in the heating surface area, such as that caused by condensate backing up into the steam space, can affect the flow of heat through the heating surface. This can cause the control system to become erratic and unstable, and processes requiring stable or accurate control can suffer with poor performance.

If heat exchangers are oversized, sufficient heating surface may remain when condensate backs up into the steam space, and reduction of thermal performance may not always occur.

However, with heat exchangers not designed to cope with the effects of waterlogging, this can lead to corrosion of the heating surface, inevitably reducing the service life of the exchanger. Waterlogging can, in some applications, be costly. Consider a waterlogging air heater frost coil. Cold air at 4 °C flowing at 3 m/s can soon freeze condensate locked in the coils, resulting in premature and unwarranted failure. Proper drainage of condensate is essential to maintain the service life of any heat exchanger and air heater.

Steam traps are devices which modulate to allow varying amounts of condensate to drain from applications under varying conditions. Float traps are steam traps designed to modulate and release condensate close to steam temperature, offering maximum plant performance, maximum plant life, and maximum return on plant investment. When stall conditions occur, and a steam trap cannot be used, an automatic pump-trap or pump and trap in combination will ensure correct condensate drainage at all times, thus maximising the thermal capability and lifetime costs of the plant. 

STEAM JACKETS

                    The most commonly used type of steam jacket consists simply of an outer cylinder surrounding the vessel, as shown in Figure 2.10.4. Steam circulates in the outer jacket and condenses on the wall of the vessel. Jacketed vessels may also be lagged or may contain an internal air space surrounding the jacket. This is to ensure that as little steam as possible condenses on the outer jacket wall, and that the heat is transferred inwards to the vessel. The heat transfer area (the vessel wall surface area), can be calculated in the same manner as with a steam coil, using Equation 2.5.3 and the overall heat  Transfer coefficients provided in Table 2.10.4.

Although steam jackets may generally be less thermally efficient than submerged coils, due to radiation losses to the surroundings, they do allow space for the vessels to be agitated so that heat transfer is promoted. The U values listed in Table 2.10.4. are for moderate nonproximity agitation.

Commonly the vessel walls are made from stainless steel or glass lined carbon steel. The glass lining will offer an additional corrosion resistant layer. The size of the steam jacket space will depend on the size of the vessel, but typically the width may be between 50 mm and 300 mm.

Table 2.10.4 Overall heat transfer coefficients for steam jackets

Process fluid or product	Wall material	U (W/m² °C)
Water	Stainless steel	850 – 1 700
	Glass-lined carbon steel	400 – 570
Aqueous solution	Stainless steel	450 – 1 140
	Glass-lined carbon steel	285 – 480
Organics	Stainless steel	285 – 850
	Glass-lined carbon steel	170 – 400
Light oil	Stainless steel	340 – 910
	Glass-lined carbon steel	230 – 425
Heavy oil	Stainless steel	57 – 285
	Glass-lined carbon steel	57 – 230

Pipes – Bare Surface Heat Loss

Heat loss vs. surface temperature.

The diagrams below indicates heat loss from uninsulated pipes with bare surfaces at still-air conditions and air temperature 70 ^oF (21 ^oC). The emissivity coefficient of the surface is 0.95.

The overall heat transfer coefficients for steam coils with medium pressure –  or hot water coils or pipes –  submerged in oil or fat:

Submerged Coils – Heat Transfer Coefficients
Type of Coil	Heat Transfer Coefficient – U –
	(W/m2 oC)	(Btu/hr ft2 oF)
Steam to Aqueous Solutions, agitated	800 – 1200	140 – 210
Steam to Aqueous Solutions, natural convection	340 – 570	60 – 100
Steam to Light Oil, natural convection	170	30

Submerged Coils – Heat Transfer Coefficients
Type of Coil	Heat Transfer Coefficient – U –
	(W/m2 oC)	(Btu/hr ft2 oF)
Steam to Heavy Oil, natural convection	85 – 115	15 – 20
Steam to Heavy Oil, agitated	140 – 310	25 – 55
Steam to Fat, natural convection	30 – 60	5 – 10
Steam to Organics, agitated	510 – 800	90 – 140
Hot Water to Oil, natural convection	34 – 140	6 – 25
Hot Water to Water, natural convection	200 – 370	35 – 65
Hot Water to Water, agitated	480 – 850	90 – 150
Heat transfer oil to Organics, agitated	140 – 280	25 – 50
Salt brine to Water, agitated	280 – 630	50 – 110
Cooling Water to Glycerine, agitated	280 – 430	50 – 75

Example – Heat Transfer from a Steam Coil

A 50 mm steam coil with outside diameter 60.3 mm(0.0603 m) and length 10 m with 1 bar absolute pressure and steam temperature 120 ^oC is submerged in an oil bath with temperature 50 ^oC.

The surface area of the coil can be calculated by multiplying the pipe circumference with pipe length as A = π (0.0603 m) (10 m) = 1.89 m²

From the table above the heat transfer coefficient is 170 W/m^2 oC for “Steam to Light Oil, Natural Convection”. The heat transfer can be calculated as

Q = (1.89 m²)((120 ^oC) – (50 ^oC))(170 W/m^2 oC)  = 22491 W  

7. Steam traps:

Steam trap selection guide – Float & Thermostatic, Inverted Bucket, Bimetal Thermostatic, Impulse and Thermodynamic Disc steam traps.

 A steam trap is a self-contained valve which automatically drains the condensate from a steam containing enclosure while remaining tight to live steam, or if necessary, allowing steam to flow at controlled or adjusted rate. Most steam traps pass non-condensable gases like air while remaining tight to live steam.

A steam trap should

• discharge condensate immediately and completely
• not leak steam
• discharge non-condensable gases like air

There are three primary categories of steam traps:

• mechanical
• thermostatic
• thermodynamic

Popular traps in these categories includes the float steam trap, the thermostatic steam trap, the inverted bucket steam trap and the thermodynamic disc steam trap. Which one is preferred depends on the application.

A steam trap prime missions is to remove condensate and air preventing escape of live steam from the distribution system. The steam trap must adapt to the application. A disc thermodynamic steam trap should never be used together with a modulating heat exchanger – and a floating ball steam trap may be overkill for draining steam pipes.

The table below can be used as a short guide for the selection of steam traps:

Steam Trap Selection Guide
Type of Steam Trap	Operation				Normal Failure Mode
	No or little load	Light Load	Normal Load	Heavy Load
Float & Thermostatic	No action	Usually continuous. May cycle.	Usually continuous. May cycle.	Continuous	Closed
Inverted Bucket	Small dribble, may leak steam with very little condensate load	May dribble	Intermittent	Continuous	Variable
Bi-metal Thermostatic	No action	Usually dribble action	May blast at high pressures	Continuous	Open
Impulse	Small dribble	Usually continuous with blast at high loads	Usually continuous with blast at high loads	Continuous	Open
Thermodynamic Disc	No action	Intermittent	Intermittent	Continuous	Open

Thermostatic Steam Traps

There are two basic designs for the thermostatic steam trap, a bimetallic and a balanced pressure design. Both designs use the difference in temperature between live steam and condensate or air to control the release of condensate and air from the steam line.

In a thermostatic bimetallic trap it is common that an oil filled element expands when heated to close a valve against a seat. It may be possible to adjust the discharge temperature of the trap – often between 60 ^oC and 100 ^oC.

This makes the thermostatic trap suited to get rid of large quantities of air and cold condensate at the start-up condition. On the other hand the thermostatic trap will have problems to adapt to the variations common in modulating heat exchangers.

• intermittent operation
• fair energy conservation
• fair resistance to wear
• good corrosion resistance
• poor resistance to hydraulic shocks (good for bi-metal traps)
• do not vent air and CO₂ at steam temperature
• good ability to vent air at very low pressure
• excellent ability to handle start up air loads
• excellent operation against back pressure
• good resistance to damage from freezing
• good ability to purge system
• excellent performance on very light loads
• delayed responsiveness to slugs of condensate
• fair ability to handle dirt
• small comparative physical size
• poor ability to handle flash steam
• open or closed at mechanical failure depending of the construction

Float Steam Traps

In the float steam trap a valve is connected to a float in such a way that a valve opens when the float rises.

The float steam trap adapts very well to varying conditions as is the best choice for modulating heat exchangers, but the float steam trap is relatively expensive and not very robust against water hammers.

• continuous operation but may cycle at high pressures
• no action at no load, continuous at full load
• good energy conservation
• good resistance to wear
• good corrosion resistance
• poor resistance to hydraulic shocks
• do not vent air and CO₂ at steam temperature
• excellent ability to vent air at very low pressure
• excellent ability to handle start up air loads
• excellent operation against back pressure
• poor resistance to damage from freezing
• fair ability to purge system
• excellent performance on very light loads
• immediate responsiveness to slugs of condensate
• poor ability to handle dirt
• large comparative physical size
• poor ability to handle flash steam
• closed at mechanical failure

Inverted Bucket Steam Trap

In an inverted bucket steam trap a leverage system multiplies the force provided by the bucket to open a valve against the pressure. The bucket is open at the bottom and robust against water hammers.

• intermittent operation – condensate drainage is continuous, discharge is intermittent
• small dribble at no load – may leak steam, intermittent at light and normal load, continuous at full load
• good energy conservation
• good resistance to wear
• good corrosion resistance
• excellent resistance to hydraulic shocks
• vents air and CO₂ at steam temperature
• poor ability to vent air at very low pressure
• fair ability to handle start up air loads
• excellent operation against back pressure
• good resistance to damage from freezing
• good ability to purge system
• good performance on very light loads – but may leak steam if there almost no condensate load for the bucket to float
• immediate responsiveness to slugs of condensate
• excellent ability to handle dirt
• large comparative physical size
• fair ability to handle flash steam
• open at mechanical failure

Thermodynamic Disc Steam Traps

The thermodynamic trap is an robust steam trap with simple operation. The trap operates by means of the dynamic effect of flash steam as it passes through the trap.

• intermittent operation
• poor energy conservation
• poor resistance to wear
• excellent corrosion resistance
• excellent resistance to hydraulic shocks
• do not vent air and CO₂ at steam temperature
• not recommended at low pressure operations
• poor ability to handle start up air loads
• poor operation against back pressure
• good resistance to damage from freezing
• excellent ability to purge system
• poor performance on very light loads
• delayed responsiveness to slugs of condensate
• poor ability to handle dirt
• small comparative physical size
• poor ability to handle flash steam
• open at mechanical failure

8. Condensate and its recovery

Condensate Pipe Lines – Sizing

Flow and pressure loss in condensate return lines – SI Units.

When sizing condensate return lines in steam distribution system – remember that the maximum condensate flow is achieved during plant start up when pipes and equipment are heated up.

A rule of thumb – the condensate load used to design condensate pipe lines should be twice the maximum production load.

Steam Condensate Pipe Lines – Sizing
Pressure loss		Condensate Flow (kg/hour)
		Pipe Size (mm, inch)
		15	20	25	32	40	50	65	80
(Pa/m)	(mbar/m)	1/2	3/4	1	1 1/4	1 1/2	2	2 1/2	3
28	0.28	90	209	380	865	1320	2554	5194	8079
29	0.29	92	214	400	878	1340	2590	5271	8196
30	0.3	93	218	403	890	1361	2631	5348	8314
33	0.33	97	226	414	930	1420	2744	5579	8677
39	0.39	107	249	469	1028	1565	3025	6142	9526
40	0.4	108	253	476	1040	1583	3062	6214	9639
43	0.43	113	263	496	1079	1646	3180	6454	10024
45	0.45	116	270	508	1107	1687	3261	6618	10297
47	0.47	119	277	521	1134	1728	3338	6777	10523
50	0.5	123	286	538	1172	1787	3447	6949	10859
53	0.53	127	296	557	1211	1846	3565	7235	11249
55	0.55	130	302	569	1238	1887	3638	7380	11476
57	0.57	133	308	580	1261	1923	3710	7525	11703
59	0.59	135	314	591	1288	1959	3783	7666	11884
61	0.61	138	320	602	1311	1996	3851	7806	12111
64	0.64	141	327	615	1338	2041	3933	7970	12383
67	0.67	146	337	634	1379	2100	4051	8210	12746
69	0.69	148	343	645	1402	2136	4119	8342	12973
71	0.71	150	348	655	1424	2168	4182	8473	13154
73	0.73	152	354	665	1447	2200	4246	8600	13336
76	0.76	155	359	675	1465	2236	4305	8723	13563
77	0.77	157	365	685	1488	2268	4368	8850	13744
78	0.78	158	366	689	1497	2282	4390	8900	13812
Steam Condensate Pipe Lines – Sizing
Pressure loss		Condensate Flow (kg/hour)
		Pipe Size (mm, inch)
		15	20	25	32	40	50	65	80
(Pa/m)	(mbar/m)	1/2	3/4	1	1 1/4	1 1/2	2	2 1/2	3
80	0.8	160	370	695	1510	2300	4427	8972	13925
82	0.82	162	375	704	1529	2331	4491	9072	14407
88	0.88	168	391	733	1590	2427	4536	9453	14651
90	0.9	170	395	740	1606	2449	4717	9548	14787
98	0.98	179	414	777	1696	2567	4944	10025	15513
100	1	180	418	785	1701	2590	4990	10115	15649
114	1.14	194	450	845	1832	2790	5366	10841	16828
118	1.18	198	457	857	1860	2830	5443	11022	17055
120	1.2	199	462	867	1880	2860	5502	11113	17282
131	1.31	209	484	907	1969	2994	5761	11657	18053
137	1.37	215	497	931	2018	3071	5906	11948	18507
140	1.4	216	502	939	2037	3103	5965	12066	18688
147	1.47	224	516	966	2096	3189	6128	12383	19187
157	1.57	231	534	1002	2168	3298	6337	12814	19822
160	1.6	234	541	1011	2195	3340	6409	12973	20049
163	1.63	237	546	1025	2218	3370	6477	13109	20276
176	1.76	246	570	1066	2309	3511	6740	13608	21092
180	1.8	249	576	1075	2331	3547	6808	13744	21319
196	1.96	261	603	1129	2440	3710	7130	14379	22317
200	2	265	611	1143	2472	3760	7221	14560	22589
212	2.12	273	629	1179	2549	3874	7430	15014	23270
216	2.16	275	634	1188	2567	3905	7493	15132	23451
220	2.2	278	641	1200	2595	3942	7570	15277	23678
229	2.29	284	655	1225	2649	4028	7729	15604	24177
235	2.35	288	664	1243	2689	4086	7843	15840	24522
240	2.4	292	672	1256	2719	4129	7927	16017	24780
245	2.45	295	679	1270	2749	4173	8010	16193	25039
255	2.55	301	694	1297	2806	4260	8176	16520	25556
260	2.6	304	701	1311	2834	4304	8260	16683	25814

Steam Condensate Pipe Lines – Sizing
Pressure loss		Condensate Flow (kg/hour)
		Pipe Size (mm, inch)
		15	20	25	32	40	50	65	80
(Pa/m)	(mbar/m)	1/2	3/4	1	1 1/4	1 1/2	2	2 1/2	3
261	2.61	305	703	1315	2844	4318	8287	16738	25900
274	2.74	313	721	1348	2917	4331	8501	17173	26554
277	2.77	315	727	1356	2935	4459	8555	17282	26717
280	2.8	317	730	1363	2948	4479	8593	17357	26839
294	2.94	325	749	1402	3025	4595	8813	17781	27533
300	3	328	757	1414	3055	4641	8900	17956	27803
310	3.1	335	771	1438	3112	4726	9063	18280	28305
314	3.14	336	775	1446	3129	4752	9074	18380	28459
320	3.2	340	784	1462	3163	4825	9204	18579	28767
327	3.27	344	792	1479	3198	4853	9299	18779	29076
333	3.33	348	801	1495	3230	4904	9408	18979	29366
340	3.4	351	809	1511	3263	4955	9516	19178	29656
342	3.42	353	813	1520	3279	4980	9571	19278	29801
353	3.53	358	825	1541	3328	5054	9707	19550	30237
359	3.59	362	834	1556	3361	5103	9798	19732	30527
360	3.6	363	835	1558	3365	5109	9809	19754	30564
372	3.72	369	850	1585	3423	5201	9979	20094	31108
376	3.76	371	854	1592	3438	5225	10024	20185	31253
380	3.8	373	859	1601	3457	5254	10081	20299	31423
392	3.92	379	873	1628	3515	5339	10251	20639	31933
400	4	383	883	1646	3554	5398	10342	20866	32274
408	4.08	388	892	1665	3592	5457	10433	21092	32614
412	4.12	390	896	1672	3608	5479	10478	21174	32750
420	4.2	394	906	1690	3646	5536	10592	21378	33090
425	4.25	396	912	1701	3670	5570	10660	21500	33294

Equipment Piping and Condensate Load

Typical maximum condensate load and equipment outlet pipe size:

Steam Condensate Pipe Lines – Equipment Outlet Pipe Sizes
Maximum Condensate Load (kg/h)	Outlet Pipe Size (mm, (in))
200	15 (1/2)
500	20 (3/4)
1000	25 (1)
2000	32 (1 1/4)
3000	40 (1 1/2)
5000	50 (2)
> 5000	65 – 100 (2 1/2 – 4)

9. Flash steam and its recovery:

Flash Steam Generation from Condensate vs. Steam Pressure – SI-units
Absolute Pressure before the condensate trap (kN/m2)	Temperature (oC)	Percent Flash Steam of Condensate (%)
		Absolute Pressure after the condensate trap (kN/m2)
		400	260	170	101.33	65	35
1500	198.3	11.3	14.0	16.4	18.9	20.4	23.2
1150	186.0	8.7	11.5	13.9	16.5	18.4	20.9
800	170.4	5.5	8.2	10.8	13.4	15.4	17.9
650	162.4	3.7	6.5	9.1	11.8	13.7	16.3
500	151.8	1.6	4.6	7.1	9.8	11.8	14.4
400	143.6		3.0	5.5	8.3	10.3	12.9
260	128.7			2.6	5.4	7.5	10.2
170	115.2				2.8	5.0	7.7
101.33	100					2.2	4.9

Example – Generated Flash Steam

A condensate system is vented to the surroundings and the pressure in the condensate system after the condensate traps is 101.33 kN/m². The absolute steam pressure before the condensate trap is 1150 kN/m².

According the table above 16.5% of the condensate will evaporate as flash steam after the trap.

Note! Without any flash recovery system the energy in the flash steam will be lost to the surroundings.

Flash Steam Energy Loss

When flash steam is generated and vented to the surroundings a considerable amount of energy is lost

If a condensate receiver in a steam distribution system is vented to the surroundings, flash steam will be generated and evacuate, and a considerable amount of energy will be lost.

The amount of flash steam generated depends on the initial and final pressure before and after condensate traps for heat exchangers. In the example – calculating generated flash steam – 0.11 kg flash steam is generated per kg condensate from a heat exchanger with 5 bar gauge pressure.

In other words 11% of the condensate is lost to the surroundings and must be replaced by new water feeded to the system.

Feeding more water requires morel heat energy and additional water treatment increasing the total operating cost for the system.

To improve system efficiency, and to reduce operating cost, it is common to limit the flash steam energy loss by using some kind of low-pressure flash recovery system where the flash steam condensates in low temperature consumers – like feeding water heat exchangers, air heaters or similar.

Basic physics behind flash steam generation

If the pressure of condensate – the saturated water at the boiling point at the actual pressure – is reduced, the heat energy in the water is reduced to a level appropriate to the final pressure. The relation pressure and the boiling temperature can be found in steam tables.

The energy – or enthalpy – made available when pressure is reduced, will evaporate a part of the water, producing flash steam.

Only a part of the condensate evaporates as flash steam. How much depends on the enthalpy in the condensate at the initial and the final pressures.

The amount of flash steam produced during the pressure reduction can be calculated:

w = (h_il – h_fl) / h_fe                              (1)

where

w = ratio of flash steam generated (kg flash steam / kg condensate)

h_il = initial condensate enthalpy (kJ/kg, Btu/lb)

h_fl = final condensate enthalpy (kJ/kg, Btu/lb)

h_fe = enthalpy of evaporation of condensate at final condition (kJ/kg, Btu/lb)

It is common that the final condition is atmospheric pressure 0 bar gauge (1 bar absolute) and temperature 100 ^oC. But closed condensate receivers with some pressure elevated condensate temperature are also used.  

Example – Flash Steam Generation

Condensate is produced inside an heat exchanger with steam pressure 5 bar gauge (6 bar absolute) on the heating surfaces. The condensate at 5 bar gauge (6 bar absolute) contains 670.9 kJ/kg of heat energy – the enthalpy – at saturation temperature 159 ^oC. The pressure in the condensate is reduced to atmospheric pressure – 0 bar gauge (1 bar absolute) when passing through the steam trap. The maximum heat energy – enthalpy – in the water (condensate) at atmospheric pressure and 100 ^oC is 419.0 kJ/kg.Evaporation energy of water at the final condition at atmospheric pressure is 2257.9 kJ/kg.

The flash steam generated can then be calculated:

w = ((670.9 kJ/kg) – (419.0 kJ/kg)) / (2257.9 kJ/kg) = 0.11 (kg flash steam / kg condensate) = 11 %

Volume of Flash Steam

The volume of condensate as flash steam is much larger than the volume of condensate as condensate.

The specific volume of condensate – or water – at 100 ^oC (212 ^oF) is 0.00104 m³/kg (0.0167 ft³/lb).  The specific volume of steam at atmospheric pressure is 1.67 m³/kg (26.8 ft³/lb).

Example – Volume of Flash Steam

By using the values from the example above we can calculate

– the volume of the condensate V_c

V_c = (1 – 0.11) (0.00104 m³/kg) = 0.00093 m³/kg

– the volume of the flash steam V_f

V_f = 0.11 (1.67 m³/kg) = 0.1837 m³/kg

This is a flash steam to condensate liquid rate of

 (0.1837 m³/kg) / (0.00093 m³/kg)   = 198

Recovery of Flash Steam

Flash steam can be used to heat consumers with temperature demand lower than 100 ^oC (212 ^oF) – like HVAC systems, hot water service systems and similar. 

Flash Steam Generation – SI-units

When condensate leaves the steam traps – flash steam is generated. Amount of flash steam generated at different pressures – kN/m².

The amount of flash steam generated depend on the steam pressure – the pressure in the condensate before it leaves the condensate trap – and the condensate pressure after the trap – the pressure in the condensate return pipe lines.

The table below indicates the ratio of flash steam generated at different pressures before and after the condensate trap.

10.       Condensate transfer pump:

High temperatures and danger of impeller cavitation is the major challenge for condensate pumping in steam systems.

Often its necessary to pump condensate generated in heat exchangers and other consumers widely distributed in a plant, back to the condensate receiver in the boiler house. A special challenge with hot condensate, which is often close to 211 ^oF (100 ^oC), is cavitation of the pump and the pump impeller.

Centrifugal pumps generates lower pressure behind the wheels and the hot condensate temporarily evaporates and expands on the back side of the vanes – before it implodes and condensates. Over time this erodes and destroys the pump impeller.

To avoid the problem there are two alternative solutions:

• Add pressure to the suction side of the pump
• Use a pressure powered pump instead of a centrifugal pump

Add Pressure to the Suction Side of the Pump

If the absolute pressure exceeds the vapor pressure at the actual temperature of the fluid entering the pump, then the Net Positive Suction Head (NPSH) is positive and its possible to avoid cavitation.

A NPSH above the manufacturers specification is important to avoid that water start boiling behind the impeller. The NPSH can be expressed as:

NSPH = (144 / ρ) (p_a – p_vp) + h_s – h_f                                        (1)

where

ρ = density of water at the appropriate temperature (lb/ft³)

p_a = absolute pressure in the condensate receiver supplying the condensate pump. This is the same as atmospheric pressure if the receiver is vented (psi)

p_vp = absolute evaporation pressure of condensate at the liquid temperature (psi)

h_s = total suction head in feet. Positive for a head above the pump and negative for a lift to the pump.

h_f = friction loss in the suction piping

According to (1) the NPSH can be increased by

• increasing the difference in the pressure in the receiver and the condensate pressure, and/or
• extend the static difference h_s by lifting the receiver or lowering the pump, and/or
• increasing the piping dimensions for minimizing the friction loss h_f in the suction pipe

If it’s not possible to increase the suction pipe and lowering the pump under the receiver it may be possible to reduce the absolute evaporation pressure in the condensate P_vp by reducing the condensate temperature with a cooling exchanger in the suction pipe.

Use a Pressure Powered Pump

A pressure powered pump use steam or air pressure to push the condensate from the receiver back to the boiler room. In principle its a simple intermittent mechanical construction working in cycle where a receiver in the pump is filled with condensate before steam or pressurized air pushes the condensate out and back to the boiler room.

External power is not required since available steam or pressurized air are used as power. There is no danger for cavitation.